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COP/RIGHT DEPOSHi 



STEAM TUEBINES 



A PRACTICAL WORK OX THE DEVELOPMENT, ADVANTAGES, AND 
DISADVANTAGES OF THE STEAM TURBINE; THE DESIGN, 
SELECTION, OPERATION, AND MAINTENANCE 
OF STEAM TURBINE AND TURBO- 
GENERATOR PLANTS 



WALTER S. LELAND, S. B. 

FORMERLY ASSISTANT PROFESSOR OF NAVAL ARCHITECTURE, 

MASSACHUSETTS INSTITUTE OF TECHNOLOGY 

AMERICAN SOCIETY OF NAVAL ARCHITECTS AND MARINE ENGINEERS 



ILLUSTRATED 



AMERICAN TECHNICAL SOCIETY 

CHICAGO 

1917 



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COPYRIGHT 1910, 1917, BY 

AMERICAN TECHNICAL SOCIETY 



COPYRIGHTED IN GREAT BRITAIN 
ALL RIGHTS RESERVED 













CU455538 



FEB 14 1917 



INTRODUCTION 

THE steam turbine is not only one of the most ancient forms 
of prime-mover, but is also one of the most recent engineering 
developments, and perhaps the most talked-of at the present 
time. Remarkable advancements in its development have been 
made during the past ten years, and this has given a great incentive 
to steam engineering, while the electrical engineer has been stim- 
ulated to work out new ideas in designing generators suitable for 
the very high speeds of rotation. The steam turbine has forged 
rapidly to the front, and in spite of the early and serious handicaps, 
in the way of steam economy, has taken its place beside the best 
reciprocating engines of the present day. The turbine possesses 
many advantages over the reciprocating engine, and its field of 
greatest influence is likely to find in the immediate future a more 
serious competitor in the gas engine than in the reciprocating engine. 

^ "With the relative merits or shortcomings of the steam turbine 
and the steam engine as prime-movers, the pages following are 
concerned to a certain extent, for modern engineering has not yet 
defined with sufficient clearness the respective fields of these agents, 
into either of which the other agent may enter with a certain degree 
of success. The rivalry between these two forms of prime-mover 
has, however, given place to the mutual recognition of the merits 
of both, and cordial co-operation in certain classes of work for which 
neither alone is fully adapted; as, for instance, the use of the low- 
pressure steam turbine with the reciprocating engine, in which a 
higher economy is effected than that obtained by the use of either 
alone. 

<J Steam turbine developments of recent years have been in the 
direction of higher economies in the production of power from 
steam, and while the ultimate limit has not yet been reached along 
this line, still the steam turbines of today are highly efficient machines, 
measuring up to the best energy economy. 

Cj The pages which follow have been written for the practical 
man, for the man less interested in the finer points of theory than 
in the results accomplished and the way they are secured by the 
most successful builders of the steam turbine. Nevertheless, the 
fundamental principles have been brought out in sufficient detail 
for intelligent mechanical design. Between the designs of dif- 
ferent builders, little attempt has been made to draw comparisons: 
but the facts have been stated with the idea that the reader mav 
form his own conclusions on points where a difference of opinion 
may arise. 




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CONTENTS 



History PAGE 

Hero t ype 5 

Watt's patent 6 

Avery and Foster type 7 

Real and Pichon type 7 

De Laval early type 10 

Parsons type 10 

Curtis type 11 

Fundamental principles 

Comparison of water and steam turbines 13 

Single-stage and multi-stage principle 14 

Action of jets and shape of impinging surfaces 15 

Nozzle 19 

Compounding 20 

Types of turbines 22 

Impulse type 23 

De Laval single-stage type 25 

Multi-stage type 27 

Reaction type 30 

Low-pressure turbines 32 

Combined with reciprocating engine 36 

Installation of turbines 41 

Performance 45 

Losses 45 

Steam consumption ; 46 

Economy of turbines 48 

Tests 50 

COMMERCIAL TURBINES 

Impulse turbines 55 

Single-stage impulse turbines 55 

De Laval turbines 56 

Riedler-Stumpf turbine 66 

Compound impulse turbines with velocity steps 68 

Curtis turbine 70 

Riedler-Stumpf turbine 70 

Terry turbine 71 

Sturtevant turbine 74 

De Laval impulse-stage turbine 76 

Westinghouse impulse turbine 77 



CONTENTS 

Impulse turbines (Continued) page 

Compound impulse turbines, pressure stages 78 

Rateau turbine 80 

Zoelly turbine 84 

Hamilton-Holzwarth turbine 88 

De Laval pressure-stage turbine 92 

Wilkinson turbine 93 

Kerr turbine 95 

Compound impulse turbine with pressure stages and velocity steps .... 99 

Curtis turbine. 101 

Riedler-Stumpf turbine Ill 

Terry turbine 112 

Buffalo Forge Spiro turbine 117 

Reaction turbines 118 

Parsons turbine 120 

Allis-Chalmers turbine 125 

Combined impulse and reaction turbines 130 

Double-flow turbine 131 

Governing. . 133 

Throttling 133 

Varying number of open nozzles 135 

Varying time of admission 136 




2,5 






STEAM TURBINES 

PART I 



Introduction. The steam turbine is one of the most recent 
engineering developments, and perhaps the most talked of, at the 
present time. During the past ten years the most marked improve- 
ments in its development have been made, and this has given a great 
impetus to engineering, especially steam engineering, although the 
very high speeds of rotation have driven the electrical engineer to 
work out new ideas in designing generators suitable for these higher 
speeds. The turbine has forged rapidly to the front, and, in spite of 
an early and serious handicap in the way of steam economy, has 
taken its place beside the best reciprocating engines of the present 
time. Many claim it to be superior in the matter of steam economy, 
but this will be discussed more fully later on. The turbine evidently 
possesses many advantages over the reciprocating engine, and, in its 
field of greatest usefulness, is likely to find in the near future a more 
severe competitor in the gas engine than in the reciprocating engine. 
For some classes of work, the steam turbine in its present state of 
development is entirely unadapted. 

The steam turbine consists essentially of nozzles or guide pas- 
sages which direct the steam onto vanes or buckets attached to the 
periphery of rotating wheels, the essential elements of which 
are shown in Fig. 1. The simplest form of turbine is perhaps 
one of the type in which a jet of steam impinges upon the buckets of 
a wheel, in much the same manner that a stream of water impinges 
upon the buckets of a Pelton water-wheel; there is, in fact, a great 
similarity between water turbines and steam turbines. The under- 
lying principles are the same in either case, but the application of 
those principles is different. Steam flowing through a properly 
designed nozzle, with 150 pounds boiler pressure on one side, and 
the usual turbine vacuum on the other, will attain a velocity of about 
4,000 feet per second, or about twice the muzzle velocity of a rifle 



2 STEAM TURBINES 

ball. Water, to attain this enormous velocity, would have to flow 
from a head of about 234,000 feet. When this is compared with 
the ordinary water head of 150 feet or less, or even with such an ex- 
ceptionally high head as 3,000 feet, which is sometimes met with in 
water powers on the Pacific Slope, a glimpse will be had of the 
magnitude of the problem confronting the steam turbine engineer. To 
put this in other words, the steam turbine designer has to deal with 
a velocity equivalent to that produced by a head of water nearly 




Fig. 1. Elements of DeLaval Turbine. 

1,500 times as. high as Niagara Falls. It will at once be seen, then, 
that the velocity of rotation of a simple turbine wheel to attain the 
best efficiency must be enormous. If this total head is to be used in 
one wheel, the peripheral speed must be nearly 2,000 feet per 
second, and at such speeds, the centrifugal force is so great that it is 
no easy matter to design a wheel that will not burst, even were there 
available some material stronger than any we now know. As it is, 
about 1,200 feet per second is considered the practical limit of 
peripheral velocity for a wheel built of the best nickel steel. 

A little mathematical calculation will show that wheels of five 
feet in diameter will revolve 4,600 times per minute to attain a velocity 
of even 1,200 feet per second at the periphery. It is the problem 



STEAM TURBINES 3 

of the steam turbine designer to reduce these speeds to more manage- 
able rates without at the same time making too great a sacrifice of 
efficiency. 

From a thermodynamic standpoint, the turbine and reciprocating 
engine are not unlike, but the force of the steam acts differently in 
them. In both, it is the heat energy of the steam that does the work. 
In the one, the steam slowly expands, exerting pressure on a piston; 
in the other, it expands in narrow passages, pushing the particles 
ahead faster and faster and thus obtaining velocity which is then 
imparted to the vanes of a rotating wheel. In the one, the steam 
acts by virtue of statical pressure; in the other, by virtue of its high 
velocity. In either case, it is the internal heat of the steam that 
causes the expansion and does the work. If heat is lost in any way, 
by condensation, radiation, etc., the work will be proportionally less. 
In the turbine, a difference in pressure from inlet to outlet acts as 
a motive force indirectly, and then only in so far as it causes a rapid 
flow of steam. 

Advantages. The well-known expression, work = force X 
space, embodies the idea that a given amount of work may be 
accomplished in a certain time by increasing the total force of the 
steam on the piston at the expense of the number of revolutions of 
the fly-wheel per minute, or vice versd. For example, a Corliss 
engine running at 180 revolutions per minute requires a mighty 
effort behind the piston to develop 1,000 horse-power, and this tre- 
mendous force demands a large cylinder, a heavy frame, and an 
immense fly-wheel. If, now, an engine were built to run at 800 
revolutions per minute, much less push behind the piston would be 
necessary to develop the same horse-power, and, therefore, the parts 
could all be made smaller, and the whole weight very much reduced. 

To go a step further, and consider the steam turbine, which 
must run at 2,000 to 3,000 revolutions per minute, it is clearly seen 
that this enormous speed reduces the mass of the parts even more. 
The heavy fly-wheel is no longer necessary, as the rotating parts are 
moving at a sufficiently high speed to acquire an immense inertia, 
and there is always a constant effort exerted on the vanes by the 
team, thereby producing an absolutely steady turning moment. 
Furthermore, the motive parts of the turbine revolve, which is in 
direct contrast to the reciprocating engine, in which the piston is 



4 STEAM TURBINES 

moving backward and forward, and the turning moment is continu- 
ally changing from a maximum to a minimum. It is clear, therefore, 
that for a given horse-power, the steam turbines produce smaller 
machines, lighter foundations, and consequently smaller power 
houses. A fair idea of the relative space occupied may be gained 
from Fig. 2. 

Again, the generator, on account of this high speed, will be 
smaller and less expensive. The turbine requires oil in its bearings 




[Fig. 2. 



Comparative Sizes of 5000-K. W. Corliss Engine and Generator and Curtis 
Turbo-Generator of Same Power. 



only; hence there is no oil to go over in the condensed steam, and 
the condensation may be used for boiler feed without any danger of 
carrying oil into the boilers. The turbine requires somewhat less 
attendance than the reciprocating engine, and the whole machine is 
compact and simple. To do its best, the turbine requires a higher 
vacuum than is ordinarily obtained for the reciprocating engine, 
and hence needs very much larger condensers, more cooling water, 
and additional air-pump capacity,, All this in a measure offsets some 



STEAM TURBINES 



of its advantages, and frequently more trouble arises from the air 
pumps and condensers than from the turbine itself. The turbine 
may of course operate at the usual vacuum with a somewhat greater 
steam consumption and a slightly lower efficiency. 

The reciprocating engine has its own advantages, and in certain 
classes of work will doubtless hold its own, but for all such apparatus 
as blowers, centrifugal pumps, generators, etc., which may be direct 
connected to a turbine, the reciprocating engine is rapidly becoming 
a thing of the past, and even for factories where belt drives are 
used, the steam turbine has been suggested. 

History. The steam turbine is not only one of the most recent 
engineering developments, but is, at the same time, perhaps, one of 
the most ancient forms of 
prime mover. In a book 
written by Hero of Alexan- 
dria, over 100 years before 
the beginning of the Chris- 
tian era, a very simple form 
of steam turbine is described. 
It consisted of a hollow sphere 
mounted on hollow bunions, 
through which the steam 
passed into the sphere. On 
opposite sides of the sphere 
were outlets consisting of 
pipes bent at right angles in 
lines tangent to the equator 
of the sphere, in such a manner that the reaction of steam escaping 
through these pipes caused the sphere to revolve on its trunions, 
in much the same way that the water escaping from the arms of 
a lawn sprinkler causes it to revolve. This turbine, which is illus- 
trated in Fig. 3, is the simplest form of the pure reaction motor. 

In 1629, Branca, an Italian, invented a turbine much like a 
miniature water wheel, which was driven by a jet of steam from a 
nozzle directed against the buckets of the wheel. This is the simplest 
form of an impulse turbine, and is illustrated by Fig. 4. 

In 1784, "Wolfgang de Kempelen designed a turbine of the lawn 
sprinkler type, similar in principle to Hero's engine, the chief dif- 




i 



Fhr. 3. 



Hero's Steam Turbine. 



6 



STEAM TURBINES 



ference being the substitution of a horizontal revolving tube for the 
hollow sphere which Hero used.' Steam, escaping from the outlets 
in opposite ends of the tube, caused it to revolve by reaction, just 
as the escaping water causes the lawn sprinkler to revolve. 

In 1784 it is 
said that James 
Watt took out a 
patent on a tur- 
bine, but, as we 
all so well know, 
devoted his gen- 
ius to the devel- 
opment of the 
reciprocating 
engine. At this 
time, both types 
of engine were 
in about the 

same crude form, and it is possible that had Watt devoted his 
energies to the turbine instead of to the reciprocating engine we 
might not have had the ordinary form of steam engine in its present 

'Knife Edge 




Branca's Impulse Turbine. 



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Js£u& 




Reaction Wheel of Avery & Foster. 



high state of efficiency, for the turbine possesses so many advantages. 
This is proved by the fact that it has at last come to the front in spite 
of the great commercial success of the reciprocating engine. 



STEAM TURBINES 



In 1831, Avery & Foster took out the first patent granted for a 
turbine by the United States Patent Office. This was on the Hero 
lines, and was really an improvement on the Wolfgang de Kempelen 
turbine of 1784. This turbine appears to be the first to attain com- 
mercial success. Several were built under the Avery patent and 
were used to run sawmills near Syracuse, N. Y. 

Steam entered a hollow 7 shaft, Fig. 5, through a stuffing box, 
passed through to the hollow arms, and escaped through plain open- 
ings in opposite ends 
of the arms. The 
speed of rotation was 
enormous, the periph- H 
ery of a 7 ft. wheel 
traveling at the rate of 
about 14 miles per 
minute. The wear 
was excessive, and 
this, combined with 
inability to get proper 
packing for the stuf- 
fing boxes, rather than 
the lack of steam 
economy, doubtless 
caused its failure, for 
the reciprocating en- 
gine of those days had 
not reached its highest fS^5^^| 
state of economy. Had 
Avery used the present Fig> 6> ReaI & Pichon Com P°und Turbine, 

expanding nozzles instead of plain openings for his steam outlets, 
his steam consumption would doubtless have been less, but the 
speed of rotation more. Diverging nozzles were used as early as 
1838, but as they were not correctly proportioned, they were a hin- 
drance rather than a help, and the idea seems to have been given up 
for a time. 

As early as 1827, a compound turbine was patented by Real 
& Pichon, the idea being to reduce the velocity of rotation by passing 




8 



STEAM TURBINES 



the steam through successive wheels G, Fig. 6, separated by disks 
B B containing outlets C to permit the passage of the steam from 
one chamber to another. H is one of the blades, F the shaft, and 
M the steam exhaust. This is the principle on which the present 
Rateau turbine works. 

The chief cause for early failure in turbine work was lack of 
comprehensive knowledge of the flow of steam. It was not until 
1840 or thereabouts that anyone seemed to get at the real facts or 
appreciate the true significance of the situation. In this year Pil- 
brow patented a machine that was a distinct advance in the right 
direction, and his patent claims show that he, at least, understood 
some of the fundamental principles. In 1842, he attempted to re- 
duce the speed of rotation by compounding, passing steam through 



Outlet 




Fig. 7. 



Outlet 
Wilson's Compound Turbine. 



successive wheels revolving in opposite directions. There are, of 
course, grave objections to such an arrangement. He later invented 
a turbine with several nozzles that could be successively cut out of 
action as the load on the turbine varied. This, in its crude form, is 
the fundamental idea of the arrangement of the nozzles used in the 
DeLaval and Curtis turbines, but Pilbrow used a converging, instead 
of a diverging nozzle, and his wheel was unlike others of the impulse 
type. 

In 1840, Wilson patented the forerunner of the Parsons type of 
turbine. He passed steam successively through rows of running and 
stationary vanes, gradually expanding it until the exhaust pressure 
was reached. A view of Wilson's invention is shown in Fig. 7; a, 6, 



STEAM TURBINES 



9 



and c, are vanes which are attached to and rotate with the drum D, 
while d, e, and / are stationary guide vanes. Steam enters at the 
left, passes through the turbine longitudinally, and exhausts at the 
right. Wilson appears to have been among the first to realize that 
the volume of steam increases as it expands to lower pressures, to 
provide for the same by increased size of passages, and, what is per- 
haps most important, to claim 
this in his patent. 

In 1858, Hartman Bros, pat- 
ented a turbine consisting of two 
revolving disks c and c' fixed to a 
shaft D, as shown in Fig. 8. 
Between them was a segment of 
stationary reversing blades dd. 
Steam entered from a nozzle F 
and was exhausted at H; G is the 
casing. This turbine embodied 
the essential element of the one- 
stage Curtis turbine of the present 
day. 

Perrigault & Farcot, about 
1870, patented a compound tur- 
bine in which the steam, as it left 
the buckets, passed through suc- 
cessive passages and again and 
again impinged upon the face of 
the same wheel. This is the prin- 
ciple adopted in the Riedler- 
Stumpf turbine and is illustrated 
by Fig. 9. Steam enters this 
turbine through the nozzle A, 
passes through the wheel buckets 
to the other side, and discharges 

into pipe B, which brings the steam around again to the inlet side. 
Here it discharges through the opening b, against another bucket of 
the same wheel, whence it is picked up by the opening x, in the pipe 
C, and so on, finally exhausting through the pipe D. Somewhat 
earlier, a turbine was invented that returned the steam in a similar 




Fig. 8. 



Hartman's Compound Impulse 
Turbine. 



10 



STEAM TURBINES 



manner, but to another set of blades on the same wheel. This idea 
has also been perfected by Riedler & Stumpf. 

Moorehouse patented, in 1877, an improvement on the type 
suggested by Real & Pichon in 1827. His chief claim was an allow- 
ance for the increased volume of expanding steam in this type of 
turbine. 

DeLaval took out a patent on a reaction turbine of the Hero 
type in 1883. It differed from the Avery turbine in detail, but not 
in principle. This turbine was extensively used for running cream 
separators, and was commercially successful, but was later aban- 
doned for the present type of DeLaval motor. 




Vnlet 



a y b l c 
Fig. 9. Compound Turbine with One Wheel. 



In 1885, Parsons took out his first turbine patent on a motor 
along the lines previously suggested by Wilson, and is responsible 
for the successful development of this type of motor. His first tur- 
bine, shown in Fig. 10, took steam in the center A, and exhausted 
at both ends through the exhaust passage E E, thus avoiding any 
end-thrust on the shaft B, At the same time, he patented his famous 
flexible bearing, now in general use. In 1888, he patented the present 
arrangement of grouping several rows of blades together increasing 
the drum diameters step by step to provide for proper expansion, at 
the same time patenting his balancing pistons, at present employed 
to relieve end-thrust. 

The expanding nozzle had been patented in 1867 for use in steam 
injectors, but it was not until 1894 that anyone patented its use in 



STEAM TURBINES 



11 



connection with a turbine. In this year, DeLaval secured this patent 
and used the nozzle in connection with his turbine, for the purpose 
of expanding the steam and getting a high velocity of jet with in- 
creased kinetic energy. 

During 1894 and 1895 there were issued a large number of 
patents, many of which have been successfully developed. Among 
them were Parsons*, Rateau's, and the first patents for the use of 
buckets of the Pel ton type. 

In 1896, Curtis patented the use of an expanding nozzle in com- 
bination with a compound wheel of the type suggested by the Hart- 




Fig. 10. Early Parsons Turbine. 



man patents in 1858. Others had used both the expanding nozzle 
and the same type of wheel, and only two years earlier patents were 
taken out for a converging nozzle with a similar wheel. From a 
study of nozzles, it will appear that the converging nozzle could be 
used economically by increasing the number of stages used in the 
expansion, but the turbine would be larger than the Curtis, and 
probably less efficient. 

Patents were issued in 1898 to Riedlei & Stumpf, whose turbine 
appears to be an improvement on the Perrigault & Farcot, patented 
about 1870. 

In 1900, the Zoelly patents were issued. This turbine in prin- 
ciple is similar to the Rateau, but different in construction. 



12 STEAM TURBINES 

The steam turbine patents issued since 1900 are altogether too 
numerous even to mention, but from them a number of commercial 
machines have been developed, and are now on the market. The 
principal commercial turbines will be described later. 

It must not be thought that this summary is at all exhaustive, 
or that even all the noteworthy turbine patents have been mentioned. 
There are hundreds of them, and it is possible here only to men- 
tion those that are the immediate forerunners of our present com- 
mercial types. This brief summary will show that the commercial 
success of the turbine has been due to a more complete knowledge 
of the properties of steam, improved details, and the possibility of 
better workmanship, rather than to the development of new principles, 
for the distinctive fundamental ideas of all of our commercial tur- 
bines had been suggested years ago. It should be borne in mind, 
however, that no fundamental principle can be successfully worked 
out unless the minutest detail is correct, and these details may, and 
in the case of the turbine did, prevent the successful carrying out of 
the early ideas. 

Fundamental Principles. The underlying principles of steam 
and water turbines are alike — a moving fluid impinges upon curved 
vanes or buckets attached to the periphery of a wheel, thus causing 
it to revolve. The vanes or buckets change the direction of the actua- 
ting fluid and absorb part of its energy, the fluid leaving the turbine 
with a comparatively low velocity. To insure reasonable economy, 
the fluid must impinge upon the vanes in a direction tangential to 
their surface at the point of impact, so as not to impart any shock and 
to avoid spattering. Further, the residual velocity at outlet should 
be as low as possible. 

In either class of turbine, rotation is caused, not by the statical 
pressure of the actuating fluid, but by the velocity which it imparts 
to the rotating turbine wheels. The kinetic energy of the fluid 

passing through the turbine is equal to — — , where W equals the 

weight of the fluid per second and V is its velocity at entrance. Evi- 
dently, the smaller W, the larger V must be to develop the same 

WV* 
power. If the fluid leaves the turbine with the velocity V a , then ■ ■ ° - 



STEAM TURBINES 13 

represents the energy not absorbed by the turbine. If V a is small, 
this wasted energy will be likewise small. 

Since the fundamental principles of the turbines are the same, 
it would seem at first sight as if steam and water turbines could be 
built on similar lines. But this is not so, because the difference in 
density and in elasticity of the two fluids requires different applications 
of those principles. In the steam turbine, not only must proper steps 
be taken to abstract the energy from the steam jet, but also to make 
that energy a maximum by providing for the proper expansion of the 
steam. 

To make more clear the differences just mentioned, it should be 
remembered that water is an inelastic fluid; that is, one having a 
constant volume under all conditions of pressure. Therefore, in 
flowing through a nozzle, if the velocity at the outlet is to be greater 
than the velocity in the pipe, the area of the outlet must be smaller 
than the cross-section of the pipe. Steam, on the other hand, is an 
elastic fluid and expands rapidly as it flows through a nozzle. If 
the increase in volume were in exact ratio to the increase in velocity, 
then, for maximum efficiency, the nozzle would be parallel-sided. 
This happens when the pressure at discharge is about 60% of the 
initial pressure. But when discharging into a low pressure, the 
volume of the steam increases more rapidly than the velocity, and 
hence, if the mouth of the nozzle is to be capable of discharging the 
same weight of steam per second as the throat (the condition for 
maximum efficiency), the cross-sectional area of the nozzle must con- 
stantly increase toward the outlet. 

Steam will expand as it passes through the turbine and if the 
passages are correctly proportioned, so that this expansion can take 
place only in one direction, that is, in the line of flow, the steam 
particles will be forced forward in a nearly uniform jet; the steam, 
by virtue of this expansion, will attain a very high velocity and the 
jet will consequently have a high kinetic energy. 

Water turbines use a relatively small head and a large quantity 
of fluid; with the steam turbines, the quantity of fluid is small, but 
the head is very large. To develop large powers with any form of 
turbine, it is necessary that a number of wheels be used. With water 
turbines, each wheel acts under full head, each using a relatively 
small quantity of water. With steam turbines, however, it is the 



14 



STEAM TURBINES 



head that must be divided into different steps; i. e., a single steam 
turbine can use the full quantity of fluid but, if desired to run at 
relatively low speeds, it can use but a portion of the total head. 
To develop 1,000 H. P. on a turbine shaft, with a head of 150 feet, 
would require approximately 4,800 pounds of fluid per second, 
depending somewhat upon the design of the wheel. With a head 
of 3,000 feet there would be required 242 pounds of fluid per 
second, and with a head of 234,000 feet, comparable with that of 
a steam turbine, the requirement would be about 3J pounds of 
fluid per second. It is thus clearly seen that the difficulty in 
developing large powers with the water turbine is that of providing 

for a sufficiently large quantity of 
fluid through the turbines, and 
with the steam turbine, that of 
handling the great velocities re- 
sulting from the enormous head. 
In all steam turbines, the steam 
is expanded in suitable nozzles 
or passages. In some the expan- 
sion is all in the nozzles; in others, 
partly in the nozzles and partly in 
the vanes or blades. In some, 
the total expansion from boiler to 
exhaust takes place in one nozzle, 
and the energy is absorbed by a 
single wheel; such a turbine is called a single-stage turbine. In 
others, the expansion in one set of nozzles is only partial, and after 
passing through one or more wheels, the steam again passes 
through another set of nozzles and set of wheels, and so on, 
until exhaust pressure is reached. This is a multi-stage turbine. 
There may be all the way from one to forty stages, or even more, 
in turbines of this type where all expansion is in the stationary vanes ; 
and in turbines where part of the expansion is in the running 
vanes, there may be 100 stages or more. Some turbines use nozzles 
for expanding the steam, and some use stationary vanes for the 
purpose, these vanes being so shaped as to provide suitable 
passage areas to permit of steam expansion. The principle is the 
same whether j nozzles or blades are used, but blades are generally 




Fig. 11. 



Apparatus for Showing Force 
of a Jet. 



STEAM TURBINES 15 

used where many stages are employed and the drop in pressure is 
small from stage to stage. 

Before taking up the actual study of steam turbines, it will be 
necessary to have a clear conception of a few elementary principles 
of mechanics. Suppose a hollow cube to be filled with some fluid 
(water or steam) at a given pressure, and to have an opening in one 
side that can readily be closed. The arrangement is such that when 
the outlet is opened, the internal pressure will remain the same. If 
the outlet is opened, the fluid will rush out, as shown in Fig. 11, and, 
if the jet is supposed to strike against a board free to move, the jet 
will exert a force upon that board tending to swing it in the direction 
of the jet. This force is called an impulse. At the same time there 
will be a tendency on the part of the cube to move in the opposite 
direction, and the force thus developed is called a reaction. It may 
be explained in this way: 

Suppose each side of the cube 
to be one foot square, the area of 
the opening, one square inch, and 
the internal pressure, 100 pounds 
per square inch. There will be 
144 X 100 pounds pressure on each 
side of the cube with the outlet 

closed, but when the One-inch Outlet Fig. 12. Jet Deflected through 90°. 

is opened, the total pressure on the 

side containing the outlet will be reduced by the pressure of 100 
pounds on the opening itself. This will leave an unbalanced force 
of 100 pounds acting in the opposite direction, which is the origin 
of the reactive force. This explanation is not strictly correct, but 
serves to give an idea of these two forces, impulse and reaction. 

Hero's turbine was a reaction turbine pure and simple, Branca's, 
an impulse turbine; but what is called a reaction turbine at the present 
time is not a simple reaction turbine in any sense, but one running 
under the combined influence of reaction and impulse. Likewise, 
the so-called impulse turbine is not a pure impulse turbine, but acts 
under the combined influence of impulse and reaction. There is 
no pure reaction turbine now on the market. The so-called impulse 
turbine being rather simpler of explanation, for the present only 
this type will be considered in the following explanations. How 




16 



STEAM TURBINES 



these principles apply to the so-called reaction turbine will be 
explained later. 

Suppose a stream of water from a nozzle to impinge upon the 
plate shown in Fig. 12, and so made that the jet is divided, and with- 
out shock departs in a direction tangential to the plate and at 90° to 
the line of impact. If the velocity of impact of the jet is V feet per 
second, its velocity in the same direction after striking the plate will 
be zero, and therefore, a definite force will be exerted on that plate, 
equal to the force necessary to impart a velocity of V feet in one second 
to the mass of water in the jet. The acceleration, therefore, will be 
V feet per second, and since force is measured by mass times 
acceleration, this force, acting on the plate, will be F = MV. If 
the plate is allowed to move in the direction of the jet with a 

velocity V v the relative velocity of 
the plate with reference to the jet 
will be V — V v and the correspond- 
ing force acting on the plate will 
be F = M (V - VJ. Since work 
is measured by the product of 
force and distance, the force acting 
through the space V x in one second, 
will do the work W = FV X = M 
(F- V t ) V t foot pounds. 

Now if the plate were shaped as 
shown in Fig. 13, so that the direction of the jet were completely 
reversed, that is, turned through 180°, there would be an additional 
pressure on the plate, due to the reaction of the jet leaving it. This, 
neglecting friction, would be equal to the original impulse, thus 
making the total force on the plate 2 F instead of F. It is quite 
evident that if the force is twice as great, the work must also be 
double, and the above expression for the work done becomes 

W = 2FV X =1X2 (V - V x ) V x 

For this reason, turbine vanes are made so as to reverse the direction 
of the jet as completely as possible. Complete reversal is not prac- 
ticable because some clearance must be allowed for the deflected jet 
to escape. This is especially true in the usual case in practice, where 
the jet impinges upon the vanes from the side. Here, the angle has 




Jet Deflected through 180 c 



STEAM TURBINES 



17 



to be such that the revolving wheel will clear both nozzle and 
deflected jet. 

If the bucket shown in Fig. 13 were held stationary, the force 
exerted by the jet would evidently be a maximum and equal to 2MV; 
but the velocity of the bucket being zero, the work, equal to 
the force multiplied by the space, 
would also be zero. If, on the 
other hand, the velocity of the 
bucket were equal to the velocity 
of the jet, the push would be zero, 
and the work again zero. Some- 
where between these limits, there 
must evidently be a velocity which 
will produce maximum results. 

V 

Suppose now, that V x = -jr ; 

then 




Fig. 14. Jet Impinging upon Curved Vane 
at an Angle with Plane of Rotation. 



W = M X 2 (V 






but 



MV 



the kinetic energy of the jet, as it issues from the 



nozzle. Therefore, if the speed of the bucket is one-half the velocity 
of the jet, we have an efficiency of 100%, neglecting losses, and 
this is, of course, the best obtainable. Therefore, the greatest effi- 
ciency is obtained when the speed of the bucket is half the jet velocity, 
provided the jet impinges upon the bucket in a direction parallel 
to the line of movement of the bucket. For other angles, the speed 
for maximum efficiency would be somewhat less. 

If a jet with the velocity V strikes the bucket at an angle a, 
as shown in Fig. 14, its velocity A B could be resolved into two com- 
ponents — one C B at right angles to the shaft, and one C A parallel 
to the shaft. The one at right angles to the shaft, commonly known 
as the velocity oj whirl, would produce a rotative impulse equal to 
Vcos a, and V v the velocity of the vane necessary for maximum 
efficiency, would be half this, or V t = J Vcos a, provided the angle 
with which the jet leaves the blade is equal to the angle of impinge- 
ment. The component A C, parallel to the shaft, would have no 



18 



STEAM TURBINES 




tendency to cause rotation, but would produce an end thrust on the 
shaft. This component is called the velocity of flow. 

Suppose a jet to impinge upon a curved vane at the angle 
shown in a Fig. 15. If the jet strikes this vane tangentially, without 
shock, the vane remaining stationary, the relative positions of the jet 
before and after impact will be as shown. Now if the vane is 

allowed to move with the velocity 
V v the relative positions of the 
vane and the nozzle will change, 
and the jet will no longer glide 
smoothly onto the vane, but will 
strike the edges, and spatter. 
To maintain the correct relative 
positions, the nozzle must either 
be allowed to follow the vane, or its 
position must be changed so that 
the direction and velocity of the 
jet will be such that it may be resolved into two components, 
one parallel with the direction of motion of the vane, and the 
other tangent to the vane. The absolute direction of the jet 
must be along the line A B f (&, Fig. 15), but its direction rela- 
tive to the moving vane will be along the line A C, and if A B 
is drawn to a scale representing the actual velocity of the jet, 
and C B laid off to the same scale to represent the velocity V x of 
the vane, then A C will represent in magnitude and direction the rel- 
ative velocity of the jet and the vane, which will be identical with 
the absolute velocity in the first case where the vane is stationary. 
Neglecting friction, the jet will leave the vane with the same relative 
velocity. Draw E F = A C and E G = C B = the velocity V v Then 
E H, which we shall call V 3 , will represent in magnitude and direc- 
tion the absolute velocity with which the jet leaves the vane. 

WV 2 

The energy in the jet before impact was 



(a) (A) 

Fig. 15. Relative Positions of Jet 

and Vane. 



WV 2 

after leaving the vane, _ 3 . 

2 9 
The energy absorbed was then 



WV 2 



WV, 



2? 



w 



*?)• 



2y 2 9 

For the best efficiency, V 3 should be small, but can never be zero 



STEAM TURBINES 19 

unless the jet angle a is zero, and the direction of the jet is reversed 
through ISO , an impracticable condition. 

Nozzles. Steam does not cause rotation in the turbine because 
of its statical pressure, but, as already stated, because of its velocity, 
a difference in pressure acting indirectly, by imparting velocity to 
the steam. It is evident, then, that in this class of motor, steam 
velocities are all-important. The steam possesses energv by virtue 
of the heat which ir contains, but to make this energy available in 
the turbine, it must be transformed into kinetic energy by the produc- 
tion of a high jet velocity. The correct shaping of the nozzle is the 
all-important factor in acquiring the requisite steam velocity, as will 
appear from the following considerations: 

It has been well established by experiment that steam at high 
pressure flowing into a space at lower pressure, through a nozzle with 
parallel sides, cannot attain a velocity exceeding 1,450 to 1,500 feet 
per second, no matter how high the initial pressure nor how low the 
pressure into which the steam discharges. This limiting velocity is 
due to the fact that at the throat of any nozzle there occurs a drop in 
the pressure of the steam to about 58% of the initial pressure, and, if 
the nozzle be a cylindrical one, this drop will remain practically con- 
stant throughout the length of the nozzle. The velocity acquired 
by virtue of this difference in pressure will therefore be about the same, 
whether the absolute pressure into which the steam is discharged is 
oS^l of the initial pressure or much less. In the latter case, the 
throat pressure cannot change until the outlet is reached, when the 
pressure drops suddenly to the pressure of the space into which the 
steam is discharging, and the steam immediately expands in all 
directions, thus dissipating its energy. The only case in which maxi- 
mum efficiency is developed with orifices and short passages with 
parallel sides is when the low pressure is greater than 58% of the high 
pressure. 

This limiting value for orifices or parallel-sided nozzles, and the 
consequent limit of steam velocity, makes it impossible to develop 
the greatest energy of the steam when expanding to low pressures 
except through a nozzle with flaring sides, in which the outlet is 
greater than the inlet. In such a nozzle the steam expansion occurs 
gradually in its flow, and is constrained to take place only in the 
direction of the flow. In this way, the velocity of the steam par- 



20 STEAM TURBINES 

tides is increased as it proceeds along its nozzle until* a tremendous 
speed has been developed, which will produce about 95% of the 
available energy. Furthermore, steam may be expanded effectively 
within the confines of such a nozzle from any high pressure to any 
lower pressure, provided the increase of areas of cross-section of the 
nozzle is proportional to the increase of specific volumes of the steam. 
In other words, with a cylindrical nozzle, a limiting steam velocity 
of 1,450 to 1,500 feet per second is possible, no matter whether the 
initial pressure of the steam be 70 pounds or 200 pounds, or whether 
the pressure into which the steam is finally expanded be 58% of the 
initial or a 28" vacuum. Of course, as the weight of the steam per 
cubic foot varies with the pressure, a greater weight of steam will 
be discharged per second at higher pressure, resulting in a somewhat 
greater kinetic energy in the steam jet. On the other hand, if the 
nozzle has flaring sides, the steam at, say, 150 pounds gauge pressure, 

will have the same pressure at the 
Press. 955 /6s. ads. Area /so. throat that is 58% of the initial 

BoUerPress\ linchMocity 1467 ft per. sec* UUUcU > UMl l *> °° /0 U1 uie «""«« 

/5oj*s s y^m^ ^— ^^ pressure, but will acquire a rapid- 

^^ ma '^'"abW'u^^ ly increasing velocity from throat 

iV/oc'ry ioW/r'per. to outlet, and, with a 28" vacuum 

Fig. 16. Properly Designed Expanding ahead of it, will leave the nozzle 

Nozzle ' with a velocity of 4,000 feet per 

second, assuming no friction in the nozzle. Fig. 16 shows a properly 
designed nozzle for expanding steam from 150 pounds boiler pressure 
to 28" vacuum. 

Compounding. It has been explained that if steam is expanded in 
a suitable diverging nozzle, nearly all the heat energy becomes available 
as kinetic energy, and that this steam, when flowing from a boiler pres- 
sure of 150 pounds to a vacuum of 28", may attain a velocity of approx- 
imately 4,000 feet per second. If the linear velocity of the buckets 
were to be approximately one-half the velocity of the jet, there 
would be grave danger that the wheel would burst from centrif- 
ugal force. A peripheral speed of 1,200 feet per second is the 
limit in practice. 2,000 feet per second would mean about 12,750 
revolutions per minute in a wheel 3 feet in diameter. This latter 
velocity would mean great delicacy in balancing and difficulty in 
providing suitable bearings, even if material could be found to 
withstand the strain. A wheel 15 feet in diameter would have to 



STEAM TURBINES 21 

revolve 2,500 revolutions per minute if the above mentioned 
peripheral velocity were to be obtained, and again the impossibility of 
construction is evident. Some means must therefore be employed to 
reduce the speed to manageable rates without unduly increasing the 
size of wheel. This may be done in a single-stage turbine by means 
of gearing, but here, if 1,200 feet is to be the maximum permissible 
peripheral velocity, and 2,000 is the theoretical velocity, there will be 
a loss of efficiency. As a matter of fact, the steam jet does not strike 
the wheel in a line at right angles to the shaft; consequently the 
velocity of whirl, as already seen, is Vcos a, and with friction allow- 
ance, this is somewhat reduced, but, even with a bucket velocity of 
1,200 feet per second, the revolutions will usually be too high. 

Turbine speeds may be satisfactorily reduced without the use 
of gearing, by what is called compounding; i. e., by dividing expansion 
into separate stages, called pressure compounding; by passing the 
steam over several wheels with guide vanes between to redirect the 
steam upon the vanes in the next wheel, called velocity compounding; 
or by a combination of these two methods. 

Suppose we start with steam at 150 pounds gauge pressure and 
expand it to 28 in. vacuum, not in one expanding nozzle but in several 
stages, so that the expansion in each would be to only about 60% of 
the next higher pressure, in which case, diverging or expanding 
nozzles would not be needed. The velocity of flow of the steam 
would be somewhat less than 1,450 feet per second at pressures above 
the atmosphere, and would decrease slightly as the pressure lowered; 
the lowest velocity, when discharging into a vacuum of 28 in., would 
be about 1,250 feet per second, but, by letting the drop in pressure be 
somewhat less than 60% in the higher stages, the velocity of flow 
could be made approximately 1,250 feet per second throughout. 
This is, of course, neglecting all losses. We could then have a steam 
speed of about 1,250 feet per second to deal with, instead of 4,000; the 

peripheral speed of the buckets would be — — ~ = 625 cos a, 

or, when a, is small, about 600 feet per second. For a wheel 5 feet 
in diameter, this would mean about 2,300 revolutions per minute, 
and the conditions arising from such a speed are much more easily 
taken care of. This reduction in speed could be accomplished in 
about ten stages. To reduce the speed to half, or 300 feet per 



22 STEAM TURBINES 

second, would require four times as many stages because the num- 
ber of stages would be equal to the square of the ratio of reduction 
of the steam velocity. Thus, the reduction from 4,000 feet per 

/4,000 V 3 

second to 1,250 feet per second will require \ 9 _ / = 3.2 =10, 

approximately. To reduce the speed of the buckets from 600 to 300 
feet per second, would evidently require four times as many, or 
40 stages. 

Suppose, now, the compounding were all in velocity stages, and 
that the expansion occurred in one nozzle. The velocity of steam 
would be nearly 4,000 feet per second, but it would have to pass 
through three sets of revolving wheels to bring the relative speed at 
each wheel to approximately 1,300 feet per second, and thus get the 
same speed of revolution as in the previous case. 

This method of compounding gives a very compact form of 
turbine and one that has many mechanical advantages; but the wheels 
have to revolve in a bath of steam which makes the friction excessive, 
and the efficiency correspondingly lower. This was the idea of the origi- 
nal Curtis patent, but was soon abandoned for the combined pressure 
and velocity turbine which is to-day the principal feature of the Curtis 
design. In this combined method, there are two or more pressure 
stages, and in present practice, not over two velocity wheels and one 
set of guide vanes to each stage.* The older Curtis had even three 
or four revolving wheels per stage with a corresponding number of 
sets of guide vanes. 

Compounding has been tried by the use of counter-running 
wheels, but with little success. If the guide vanes were on wheels 
free to turn, they would run in a direction opposite to the others, 
and if the relative peripheral velocity of the two were, say, 1,200 feet 
per second, it would mean that for each wheel the absolute velocity 
would be half this, or 600 feet per second. The difficulties of build- 
ding and operating such a machine are considerable. 

Types of Turbines. There are two main groups into which 
steam turbines are usually divided, one known as the impulse, and 
the other as the reaction type. It has become the general practice 

*This does not apply to marine practice, the peculiar conditions of which 
warrant the use of a larger number of velocity wheels per stage. Small 
Curtis turbines and some special machines have three velocity stages. 



Van* 



STEAM TURBINES 23 

to classify turbines under one or the other of these two general heads; 
but, as a matter of fact, every commercial turbine of the present time 
really develops its power under the combined influence of action 
and reaction. Yet there is a distinct difference between the expansion 
of steam in these two types, as, for instance, in the DeLaval and 
in the Parsons turbine. The use of the terms impulse and 
reaction in reference to turbines is undoubtedly unfortunate, but 
since their use has become practically universal, it is necessary to 
understand the significance of their application. 

In the so-called impulse turbine, the steam, expanding in a 
nozzle or other suitable passage, thus attains a high velocity, and 
impinges upon the vanes of a rotating wheel. The steam, in passing 
through the wheel, gives 
up a part of its kinetic 
energy to the revolving 
vanes, and leaves the 
wheel at a lower velocity, 
but at the same pressure 
at which it left the noz- 
zle. In the so-called re- 
action type, the steam 'l[[jH 
enters the turbine at ^ (t) 

boiler pressure, passes Fig. 17 Two Arrangement of Jet and Vane, (a) Pure 
,. . . , Action. (6) Action and Reaction. 

through guide passages 

onto the vanes of a rotating wheel, and little by little expands as it 
passes through these vanes to subsequent guide passages and other 
vanes, the pressure gradually becoming lower; the velocity which is 
gained by the expansion in the guide passages and revolving vanes is 
practically all imparted to the rotating drum. The pressure is less 
on the one side of the vane than on the other, while, in the impulse 
type, the pressure is the same on both sides of the vane. 

The engines of Hero, Wolfgang De Kempelen, and Avery were all 
purely reaction types, but the Parsons acts by impulse as well as 
reaction, and the Curtis and DeLaval, by reaction as well as by im- 
pulse. To make this clear, consider Fig. 17 (a), which shows a vane 
and jet. The vane is so shaped that the jet leaves it at right 
angles to the direction of impact. Here is a case of pure action, so 
far as any force tending to move the vane in a direction parallel to the 




24 



STEAM TURBINES 



direction of the jet is concerned. There is, to be sure, a reaction of 
the jet, but this reacting force is along a line A B at right angles to 
the desired line of motion, and if the vane shown in this figure were 
attached to the periphery of a wheel free to revolve, this force of 

reaction would cause only an end- 
thrust on the shaft, in no way 
augmenting the force of rotation. 
As previously shown on Page 
16, to obtain the best efficiency 
the jet must be deflected through 
an angle of 180°. If the jet leaves 
the wheel at any less angle than 
90°, for instance, angle B AC m 
Fig. 17 (b), there is a reactive 
force along the line AC, which can 
be resolved into two components 
— one, A B t tending to cause 
rotation, the other, B C, causing 
an end thrust. A turbine thus 
constructed, although called an 
impulse turbine, evidently derives 
an impelling force from this reac- 
tion. A pure impulse turbine, 
permitting no reaction of the jet, 
would have a theoretical maxi- 
mum efficiency of only 50%. 
When the jet is turned through 
an angle of 180°, the reaction 
becomes equal to the impulse. 
The reaction is equal to the im- 
pulse in any case, when the angle 
at which the jet impinges upon 
the vane is equal to the angle of 
deflection measured from a plane 
through the center of the rotating wheel at right angles to the shaft. 
In this type of turbine, all the expansion takes place in the nozzles 
or guide passages, none at all in the revolving vanes. 

In the so-called reaction turbines, the expansion takes place 




Boil erPress. 



Throat Pre&i\ 

95.7~*absJ* 
Throat \fei \ , 
jer.secj , 



f500p< 



Vel. t'nSteam 



,Lost ve1.t oCon&. 



/n!et'\ 'Throat >*Outlet^ 



£jrh.28£ vac. 
C Base line 



•Sfeampipe ~**+/Vozz?e**rVar}e$*£jrhau&t pipe 
^Clearance 



Fh 



. 18. Typical Features of Single-Stage 
Impulse Turbine, with Relations of 
Steam Pressures and 
Velocities. 



STEAM TURBINES 25 

in the revolving vanes as well as in the guide passages, and the 
vanes and guides are placed so as to give a constantly increasing area 
of passage to allow for the increasing volume of the steam as it 
expands. As the steam expands in the guide passages, it acquires 
velocity and impinges upon the running vanes, thus giving a decided 
impulse to them, and as it again expands in the running vanes, the 
reaction produces a further impelling force. 

The distinguishing feature, then, between these two distinct 
types of turbine is not to be found in the impulse or reaction of the 
steam at all, for both types, as we have seen, act by virtue of both 
forces; but the distinction lies in whether the expansion of the steam 
takes place fully in a set of nozzles or guide passages with no expan- 
sion in the moving vanes, or whether the steam expansion takes place 
partly in the nozzles and partly in the revolving vanes. A turbine 
might be so arranged that the expansion would take place entirely in 
the moving vanes, the guide passages acting merely to change the 
direction of the steam, but as yet no commercial turbine has been 
built on these lines. 

There are several distinct subdivisions of the two main types 
of turbine. The simplest form is undoubtedly of the DeLaval type, 
which consists of several diverging nozzles, expanding the steam 
from boiler pressure to exhaust pressure, and directing the steam jets 
onto the vanes of a single wheel. We have seen that the enormous 
velocity of 4,000 feet per second will be attained in expanding from 
150 pounds boiler pressure to 28 J inches vacuum. The speed of 
revolution must be very high and, although the velocity is greatly 
lowered as the steam passes through the wheel, it will leave the 
wheel with a considerable residual velocity which represents, of 
course, so much lost energy. Fig. 18 illustrates the typical features 
of this style of machine, the curves showing the relation of its 
steam velocities and pressures. It will be noticed that the steam 
pressure is a maximum, and equal to boiler pressure, at the inlet to 
the nozzle, and will reach the condenser or exhaust pressure at the 
nozzle outlet, as it impinges upon the vanes of the wheel. Clear- 
ance, in this type of turbine, is of small consequence, for the wheel 
revolves in steam of a uniform pressure, and there can, therefore, 
be no leakage of steam without work being done. As there is 
but one wheel revolving in the bath of steam, the friction would 



26 



STEAM TURBINES 



not be very great, were it not that the friction increases very rapidly 
with the speed, and in this single-wheel type, the speed of the steam is 
very high. The chief loss will be due to the relatively high velocity 
of the exhaust steam, and to the friction of the bearings on account of 
the high rotative speed. To reduce these speeds of rotation to man- 



Vanes 



Vanes 



Vanes 




Shaft 



Met' Outlet* , 



Exhaust Press. 



Steam Pipe +Ntizzle Wanes* Wozzle "Vanes* ^Nozzle Wanes* ^Nozzle 'Wanes*- £xh oust Pipe 

^Thmn* ^Thr-nrr* *Thrnrrt 'Throat 



Throat 



Throat 



Throat 



Fig. 19. 



Features of Multi-Stage Impulse Type, Showing Relations of 
Steam Pressures and Velocities. 



ageable rates, gearing must be used, causing a further frictional 
loss, or the diameter of the wheel must be abnormally great. 

The velocity of the steam at the entrance to the nozzle is that 
due merely to its flow through the pipe. At the throat of the nozzle, 
the velocity, as we have previously seen, will be something under 
1,500 feet per second, and at the mouth of the nozzle, if it is properly 



STEAM TURBINES 27 

designed, the velocity will approximate 4,000 feet per second, assum- 
ing a boiler pressure of 165 pounds absolute and 28 J inches of 
vacuum. This high velocity will not be maintained, however, as 
the steam passes through the revolving vanes, but, at the condenser, 
will have dropped to a value depending upon the amount of energy 
absorbed from the steam during its passage through the vanes of the 
wheel. 

If wheels were arranged in successive chambers, so that the 
steam could be expanded in several steps instead of in one, we should 
have the essential elements of the Rateau type of turbine. Fig. 19 
shows diagramatically the essential features of this type of turbine, 
and the relation of velocities and pressures, as before. Each wheel 
rotates in an independent chamber separated from the next by a 
diaphragm provided with suitable expanding passages, so that the 
steam, in passing from the first chamber to the second, will be under 
conditions similar to those obtaining when passing from the boiler 
into the first chamber, and again may attain a maximum velocity. 

In a four-stage turbine of this sort, the pressure, as shown in the 
curve in Fig. 19, should be a maximum (boiler pressure) at the inlet 
to the first nozzle. At the throat of the nozzle, it should be approx- 
imately 58% of the initial pressure. During its passage through the first 
chamber, the steam pressure would be constant, and it would again 
drop in a similar manner, in passing through the nozzles between the 
first and the second chamber, the velocity rising with each drop in 
pressure. With a four-stage turbine, the drop in pressure would be 
such that one-fourth of the total available heat units would be 
available in each chamber. The drop in pressure from chamber to 
chamber would therefore not be uniform, for a given pressure 
change represents more heat units in the lower than in the higher 
ranges of pressure. The velocity at the inlet to the first nozzle would 
again be merely the velocity of flow through the steam pipe; at the 
throat of the nozzle, approximately 1,500 feet per second, and at the 
outlet to the nozzle, where the steam impinges upon the vanes of the 
first wheel, approximately 2,000 feet per second. This velocity will 
drop as the steam passes through the wheel, rising again on its pas- 
sage through the next nozzle, dropping again in the next wheel, and 
so on, the residual velocity as the steam leaves the last wheel being 
probably less than in the previous case. 



28 



STEAM TURBINES 



Moving 
■ .Vanes 



Turbines are built on this principle by a number of manufac- 
turers, the Rateau being the best known of this type. This particular 
turbine has usually a large number of chambers, frequently 30 to 40, 
and the drop in pressure from chamber to chamber is consequently 

very small, so small in 
fact that expanding noz- 
zles are not necessary. 
This type of turbine is 
subject to leakage at A B, 
Fig. 19, where the shaft 
passes through the sta- 
tionary diaphragm and 
requires special packing. 
This packing becomes 
evidently inaccessible in 
a multi-stage turbine. 
A simple method of 



Nozzles 




compounding, but one 
not likely to produce as 
economical results, is 
that shown diagramatic- 
ally in Fig. 20, its vari- 
ations of pressure and 
volume being shown in 
the curve. In this tur- 
bine, steam is expanded 
in a properly designed 
diverging nozzle, from 
boiler pressure to exhaust 
pressure, and impinges 
successively upon the 
vanes of rotating wheels. 
Between these wheels 
are stationary guide vanes curved in the opposite direction, so that, 
as the steam leaves the first set of vanes, it is redirected by these 
guides upon the next set, and so on. The boiler pressure is exactly 
similar to the boiler pressure shown in Fig. 18; the velocity of the 
steam as it leaves the nozzle is also the same. This velocity drops 



Station* 



•SteamPipe-\<\NozzlettMoi'ing>yanj *ffio^ing\*aru \?Movinq*t 
V n Vanes ''Vanes 1 Vanes \ Vanes Vanes 11 



Base Line 
Ejc Pipe 



Fig. 20. 



Features of Impulse Turbine Compounded 
by Velocity Steps Only. 



STEAM TURBINE^ 



29 



Moving 
[Vanes 



somewhat as the steam passes through the first set of running vanes, 
remains constant as it passes through the first set of guide vanes, 
again drops in the next set of running vanes, and again becomes con- 
stant in the guide vanes, and so on; the velocity of the steam jet is 
gradually lessened as it passes through wheel after wheel. The drop 
in velocity in the steam in its passage through any one set of vanes 
will, neglecting losses, be 
approximately equal to 
the total velocity divided 
by the number of sets 
of running wheels. In 
this type of turbine, since 
the velocity of the steam 
is gradually decreased, it 
is evident that if the same 
quantity' of steam is to 
flow through successive 
wheels in the same inter- 
val of time, the passages 
must gradually increase 
in size. The velocity re- 
maining constant in the 
guide vanes,they may pro- 
vide passages of uniform 
section, as shown in Fig. 
20, each set of passages, 
however, being larger 
than the preceding set. 
The principle just 
described was the origi- 
nal idea claimed in the 
early Curtis patent, but was subsequently given up for the im- 
proved arrangement shown in Fig. 21. This arrangement differs 
from the other, in that, instead of fully expanding the steam in one 
nozzle or set of nozzles from boiler to exhaust pressure, the expan- 
sion is divided into two or more stages. This turbine contains cham- 
bers, just as the Rateau type does, the difference being that in the 
Curtis, each chamber contains two sets of running wheels and one set 




Boiler Press 



,/iroaf Press. 



Vel in Stearmj Pipe 

Inlet' Tnroat^Outtet,^^. ,i inlet' > 
:carr. Pipe --cNopzle i yones y a „ es yanes, 'Nozzle ' Wanes Varies Vanei 



Fig. 21. Features of Turbine Compounding by 
Pressure Stages and Velocity Steps. 



30 



STEAM TURBINES 



of guide vanes, while the Rateau chamber contains only one wheel 
and no guide vanes. Turbines of the Curtis type have from two to 
seven pressure stages, but at the present time, no more than two 
sets of running vanes are used in each chamber*, although formerly, 
more sets of running vanes were used. The relation of pressures to 
velocities shown in Fig. 21 will be evident from the previous ex- 
planations. 

In the reaction tur- 
bines, of which Parsons' 
is the best known ex- 
ample, the steam, as 
already stated, gradually 
expands in passing from 
boiler to condenser pres- 
sure. The velocity rises 
in the first set of station- 
ary vanes, and drops as 
the steam does work in 
the first set of running 
vanes. The velocity rises 
again in the next set of 
stationary vanes, drops 
in the moving vanes, and 
so on. Fig. 22 shows the 
essential features of this 
turbine and the relation 
of pressures and volumes. 
The stationary guide 
vanes act just like small 
nozzles, and allow the 
steam to expand and acquire velocity. The moving vanes also allow 
the steam to expand, and the reaction of this expansion gives an 
added impulse to the rotating wheel. 

It is not intended that the foregoing shall be a description of 
any turbine, but merely a description of the distinct and elementary 
features of the action of steam in various types of turbine. 



Pipe, 




BoilerPreas. 



__ _. Guida.ffov/ng,, Guide' ftovind.'&uide , Roving \ Guide" Mo* 

Steam Pipe v ant s Wanes j *Vanes\ front J fanes\ y<rnes j^/wl m 

Fig. 22. Features of Reaction Type 



vZ%\£*-*» 



*See foot-note, Page 22. 



STEAM TURBINES 31 

In the one-stage, compound-velocity turbine, the steam leaves the 
nozzle at exhaust pressure with a high velocity. If the expansion has 
been complete, as intended, the pressure remains constant as the steam 
passes through the turbine, and there is no tendency to leakage. 
The clearances between the blade tips and the casing can be made 
as large as convenient, for it requires a difference of pressure to cause 
steam leakage. If running on vacuum, there would be a tendency 
for air to leak in around the shaft, and consequently this would need 
to be well packed. 

Here would seem to be a happy solution to the problem of steam 
leakage, at the same time producing a most compact form of turbine; 
but, unfortunately, a considerable loss is brought about by steam 
flowing past the surfaces of both moving and guide vanes and by the 
large amount of friction, due to the rotation of the many wheels 
through the steam which fills the turbine. A further serious disad- 
vantage is that an equal amount of work cannot be done in each set 
of vanes if the entrance and exit angles of the vanes are made 
equal, as is usually the case. For example, suppose a 4,000 foot 
steam velocity to be reduced in four wheels, each wheel absorbing 
1,000 ft. per sec. Then, if V v V v V v and V A represent the respective 
velocities at the entrance of each wheel, the available energy is 

WV 2 WV 2 W 
for the first wheel, -^ - -^T = Tg X 7 > 000 > 000 '> 

WV 2 WV 2 w 
for the second, -y-2- ^ = y" x 5,000,000; 

W 

for the third,} — X 3,000,000; 

W 
and for the fourth, -jr- X 1,000,000. 
2 9 

This difficulty will be remedied by increasing the number of 
pressure stages, and decreasing the number of wheels in each stage 
to a minimum. With a large number of velocity compound wheels, 
the work done by the last wheel would be so small that the frictional 
losses would be too large to make it at all economical. For example, 
with six wheels, the last wheel would develop only 9% of the power 
developed in the first. In turbines of this type, by a suitable 



32 STEAM TURBINES 

design of the nozzles and entrance and exit angles of the vanes, the 
same amount of steam energy may be abstracted in each pressure stage. 

The leakage in this type would be relatively small, only what 
would pass from stage to stage. This would be comparatively small, 
because the steam could escape only through the opening where the 
shaft passes through the diaphragm (A B Fig. 19) that separates the 
two chambers, and with small clearances this could not be large. 
With a large number of stages, as in the Rateau turbine, leakage in 
the high pressure end is not all lost, for it has an opportunity to 
work in the lower stages. 

In the reaction turbine, leakage of steam is a most important 
factor. As the pressure on the two sides of the vane is different, 
there is a tendency for the steam to escape between the tips of the 
vanes and the outer casing A B, Fig. 22, also between the ends of the 
guide blades and the rotor C D, Fig. 22. As the rotors are of large 
diameter, a large area is offered for leakage, unless the clearances 
are kept very small. Here, the steam leaking from the higher pres- 
sures, will of course do work on the lower pressure, but at a less 
efficiency, just as in the Rateau type. The successful turbine of this 
type requires great nicety of workmanship in order that the clearances 
may be adjusted to a minimum. 

Low=Pressure Turbines. The greatest drawback to improve- 
ment in any existing engine plant, or, in fact, in any mechanical in- 
stallation, has always been the fact that the equipment already in- 
stalled must be discarded, often thrown into the scrap pile, while still 
in fairly good condition and capable of doing a considerable amount 
of work. In the early installation of steam turbines, this was often 
done, and in order to increase the capacity of the central station, good 
reciprocating engines were often thrown out and turbines put in their 
places. It was, however, soon discovered that this, in many cases, 
was unnecessary, and that the desired increase in power could be had 
Vy simply using low-pressure turbines in connection with the exist- 
ing reciprocating engines. The low-pressure turbine takes the steam 
exhausted by the engine, slightly above the atmospheric pressure, 
and expands it to a lower vacuum than could be economically done 
in the engine. 

While the reciprocating engine is highly efficient for utilizing 
the available energy of steam between boiler and atmospheric pres- 



STEAM TURBINES 33 

sure, it is relatively inefficient for utilizing the energy of steam in the 
lower ranges of pressure, especially at pressures below 20 in. vacuum. 
The steam turbine, on the other hand, utilizes the available 
energy of steam in the lower more effectively than in the higher 
ranges of pressure. Since there is about as much available energy in 
steam below the atmospheric line as there is in steam above it, there 
is every reason to believe that this combination of engine and turbine 
will be a most efficient one. In order that the possibilities and limi- 
tations may be fully stated, however, it will be necessary to investi- 
gate some of the characteristics of steam expansion. 

A single cylinder engine with cut-off at, say, one-third stroke, will 
expand the steam to three times its initial volume, and if it takes steam 
at 150 pounds gauge pressure, the volume of each pound of that steam 
before expansion will be approximately 2.75 cubic feet. Now, if this 
is expanded to three times its initial volume, every pound of steam 
entering the cylinder will, at exhaust, occupy 3 X 2.75 = 8.25 cubic 
feet. If the expansion has been adiabatic, that is, without the gain 
or loss of heat, this pound of steam will occupy 8.25 cubic feet of 
space when the pressure has reached 32 pounds by the gauge, and, 
under the above conditions, an engine would release at this pressure — 
a manifest waste. 

With one-fifth cut-off and five expansions, the final volume of one 
pound would be 5 X 2.75 = 13.75 cubic feet, and this volume would be 
reached at about 11.7 pounds gauge pressure. Fig. 23 will illustrate 
this. The line b c d e is a curve representing the relation of pressures 
and volumes of steam, as it expands adiabatically from 150 pounds 
gauge pressure to the atmosphere and beyond the atmosphere into 
partial vacuum. The total available work in the steam above at- 
mospheric pressure, would be represented by the area of the diagram 
ab eh. The greatest possible work that could be done in the cylin- 
der, cutting off at one-third stroke and exhausting at atmospheric 
pressure, would be the area ab c g h, which shows that a considerable 
amount of the energy is lost. Even cutting off at one-fifth stroke, the 
work represented by the area d e / is lost. To carry the expansion of 
steam in a single-cylinder engine even to one pound above the 
atmosphere, the boiler pressure must be greatly reduced, or the amount 
of expansion increased materially. If this engine were made con- 
densing, m k would represent the back-pressure line, and while the 



34 STEAM TURBINES 

total available energy would be increased by the area h e mk, the 
work in the cylinder at one-fifth cut-off would be increased only by the 
area hfjk,a, very small part of the whole. In such case, the gain 
would probably not pay for the cost of maintaining the vacuum. 

In a compound two-cylinder engine taking steam at 150 pounds 
gauge, the ratio of high- to low-pressure cylinder volumes would be 
not over 1 to 5, and with cut-off on the high-pressure cylinder at one- 
third stroke, there would be room for not over 15 expansions; that is, 
the volume of steam at the end of the low-pressure stroke would be 
not over 15 times the volume of the steam admitted. Now, if one 
pound of steam at 150 pounds gauge pressure were expanded to 15 
volumes, the result would be 15 X 2.75 = 41.25 cu. ft. One pound 
of steam thus expanded from 150 pounds pressure will occupy 41.25 

a b 150 lbs. Gauge 



32 /bs Gauae 



fl-7 lbs. Gauqe 
^~- — - eAtmosph eric Line 



Fig. 23. Relation of Pressure and Volume in Non-Condensing Reciprocating Engine. 

cubic feet when the pressure has reached approximately 7.5 pounds 
absolute, which would correspond to a vacuum of approximately 15 
in. In other words, neglecting the condensation and other losses in 
the cylinder, the ordinary compound engine with Corliss gear (an 
engine in every way first class), cutting off at one-third stroke, cannot 
expand steam at 150 pounds boiler pressure lower than to 15 in. vacu- 
um. Any increase of vacuum beyond this point tends only to reduce 
the back pressure on the piston, and the gain in work is slight, per- 
haps not enough to pay for the additional work on the air pump, 
increased size of condenser, and additional circulating water. 

Fig. 24 shows, as before, the adiabatic expansion from 150 
pounds gauge pressure. If a b represents one volume, h f would 
represent fifteen, h f will be the back pressure line at 15 in. vacuum, 
the maximum theoretical work done in the cylinder will be the area 



STEAM TURBINES 35 

ah dj h, and the work lost will be the area h f m k. Increasing the 
vacuum below 15 in. gives only a little gain, represented by the area 
h j g k, although more than in the previous case. 

A triple-expansion engine will permit of about twenty expansions; 
that is, the low-pressure cylinder will contain about twenty times 
the volume displaced by the piston at cut-off in the high-pressure 
cylinder. In such an engine, the final volume of one pound of steam 
expanding from the previous pressure will be 55 cubic feet, and the 
pressure corresponding to this volume would be 5.5 pounds absolute, 
equal to about 19 in. vacuum. A condenser giving 24 in. vacuum 
would allow just about difference enough to give a ready flow of steam 



Fig. 24. Relation of Pressure and Volume in Condensing Reciprocating Engine. 

from the engine to the condenser. If a greater vacuum is to be used 
to advantage, the number of expansions must be increased. Even 
here, increasing the vacuum beyond 19 in. gives relatively little gain 
in the engine. To expand steam from 150 pounds gauge pressure 
to 28.5 in. vacuum would require a final volume of 338 cubic feet for 
each pound of steam admitted to the cylinder, and since one pound at 
initial pressure occupies 2.75 cubic feet, the steam would have to ex- 
pand 338 -2- 2.75 = 123 times, approximately. The utter impossibil- 
ity of such expansion in the triple expansion engine will be evident 
from the following consideration: 

If a triple-expansion engine were to expand the steam to this 
pressure, with cut-off at one-third stroke, the low-pressure cylinder 
would have a volume 123 -s- 3 = 41 times that of the high-pressure 
cylinder, and its diameter would be to the diameter of the high, 
as 1 is to the square root of 41, or about 6.5. This ratio is not 
far from three times that found in general practice for such an engine, 



36 STEAM TURBINES 

and about four times that for a compound engine. Assuming that 
the low-pressure cylinders are now as large as they can conveniently 
be made, the complete expansion above outlined would require, 
in the triple-expansion engine, three low-pressure cylinders of the 
present size. Radiation loss and friction could easily overcome the 
theoretical gain; to say nothing of the prohibitive cost and weight of 
the engine. 

Consider the diagram in Fig. 25, which shows, as before, the 
adiabatic expansion between 150 pounds gauge and 28.5 in. vacuum. 
The black area represents the available work due to the complete 
expansion of the steam, in excess of that available in the triple-ex- 
pansion engine, running under 28.5 in. vacuum. This lost energy is 

n l&S ibs. absolute 



Atmoepheric Line 



20 vac. _ 

£5" vac. -S tojiiiuiua z^mma^^^^—^mm— i . . . ■ ■ — 1 

i *-20 Volumes # ^ . , J 

k 124 Volume9 ■ - ■ h 



Fig. 25. Energy in Steam not Available for Reciprocating Engine. 

about 25% of the total energy available in the steam, or about 35% 
of the energy available for use in the reciprocating engine with 28.5 in. 
back pressure. Under the ordinary conditions of 25 in. back pressure, 
the black area would be augmented by the crosshatched area, mak- 
ing the lost energy about 40% instead of the 35% above. All of this 
energy is lost by the triple-expansion engine, but can be utilized by 
the turbine. Low vacuums cause large initial condensation in recipro- 
cating engines, but do not have any disadvantageous effect on the 
turbine. 

The low-pressure turbine can be advantageously used in 
connection with any reciprocating engine, and their combination 
will always afford a considerable improvement in economy, and 
increase the power without increasing the size of the boiler plant. 
It often happens that engines are operated non-condensing because of 



STEAM TURBINES 37 

the expense of cooling water, and as we have already shown, the rel- 
atively small gain would not pay for additional complications and 
expense, especially if cooling towers have to be provided. The low- 
pressure turbine, however, will provide enough additional power to 
pay for the installation of proper equipment. There are already in 
existence, plants where low-pressure turbines have been installed in 
connection with engines previously used as non-condensing, and the 
output has been practically doubled without increased cost for fuel. 

It is readily seen from the previous discussion, that even in a 
plant in which the engines are operated as condensing engines, a 
considerable gain can be effected by installing a low-pressure tur- 
bine, even though using the same condenser facilities as before. 
In some ways, it is much easier to maintain a high vacuum in such 
a combination, because the turbine will take the steam at slightly 
above the atmospheric pressure, and thus prevent a considerable 
amount of air leakage, which always takes place through the the stuf- 
fing-boxes of a low-pressure reciprocating engine. 

If saturated steam expands adiabatically from 150 pounds 
gauge to a pressure of 28.5 in. vacuum, practically half the 
available energy is developed between the initial pressure and one 
pound above the atmosphere, and the other half below the latter 
pressure. It might be said, in explanation, that the work of expan- 
sion can be considered as equal to the pressure times the volume; but 
it is, perhaps, not often realized that the volume of steam will nearly 
double in expanding from 26 in. vacuum to 28 in., and that, 
therefore, the available energy is great, although the pressure is low. 
In most condensing engines, the gain over non-condensing conditions, 
as determined by actual experiment, does not exceed 30%, even under 
favorable conditions of steady load. Under average conditions, the 
gain drops to 25%, and under overload conditions, to a still lower 
point. In general, a condensing reciprocating engine, if run non- 
condensing, will carry about 70% of its maximum load, exhaust- 
ing at, say, two pounds above atmospheric pressure, and, if the 
steam from such an engine be exhausted into a low-pressure turbine 
with proper condensing facilities, the latter will develop nearly as much 
work as was developed by the engine itself, and there will result 
from the two about 140% of the work which might be expected 
from the reciprocating engine alone, if run condensing. It is inter- 



38 



STEAM TURBINES 



esting to note that the discussion of Fig. 25 seems to show a 
possible theoretical gain of about 40% over the engine condensing 
at 25 in. vacuum, provided the turbine is run at 28.5 in. vacuum. 

Fig. 26 shows a study of the possibilities in connection with a 
Rice-Sargent engine which has been operated for some years in the 
plant of the General Electric Company at Schenectady, N. Y. This 
unit operates a 250-v. direct-current generator, and ordinarily runs 
with a load of 1,200 kw. 









































Tests of a 


22x44x42 Rice Sargent engine 


34 






































With D.C.ZSOVolt Generator 
Compared with 
Results obtainable from same engine in 
connection with low pressure A. C • 
Turbine unit. 

Points marked O, X, ®. correspond to 
equal steam flows. 
























P 














32 






















/ 






















^hr. 


. 






£\P 
























°l 


m-condj^ 


& 




















$28 


































































P 










































































4 










































































n 


/ 








































































$ 


















































& 
























j 










































































f e 


? 


r 




















































i** 


















t 


f 






































































































































$20 


























































































































































/a 


























































































































































16 


































71 


Chgine ond low press 
_, 28'Vacuum " ^ 


jn 


' Turbb 


ie 














/« 




































Til Ml" 



























70O 800 t/OO 1300 /500 1700 /900 2100 2300 2500 

K W.Output 

Fig. 26. Curves Showing Economy of Engine with Low-Pressure Turbine. 



*The tests were made accurately by weighing condensed steam, the effect 
of vacuum being determined by holding the steam flow constant, and chang- 
ing the vacuum. The curves show performance under condensing and non- 
condensing conditions, and also show what could be accomplished by this 
engine in combination with a good low-pressure turbine. The rates of gain 
here shown will seem extraordinary, but they are fairly representative of the 
possibilities in the average condensing engine plant,. 

Referring to the curve-sheet, note that the upper curve represents the 
engine operating non-condensing at 810 kilowatts, the steam consumption 
being 30.6 pounds per kilowatt. With the load increased to 1,065 kilowatts, 
the steam consumption is still 30.6 per kilowatt and with the load increased 
to 1,265 kilowatts, the steam consumption is 33.6 [pounds. Operating under 
these conditions, 1,265 kilowatts is practically the maximum capacity of 
the unit. 

Now, operating condensing with a capacity of 1,140 kilowatts, the steam 
consumption is 22 pounds per kilowatt; at 1,320 kilowatts, the steam consump- 
tion is 24.6 pounds per kilowatt; and operating at 1,470 kilowatts, the steam 
consumption is 28.8 pounds per kilowatt. Note, however, that the maximum 
capacity of the unit has been increased from 1,265 to 1,470 kilowatts. 

*From a paper by Chas. B. Burleigh, on the "Low-Pressure Steam Turbine." 



STEAM TURBINES 39 

Now, by the assistance of the low-pressure turbine, vacuum conditions 
remaining the same, the steam consumption at 1,550 kilowatts is 15.6 pounds 
per kilowatt; at 2,020 kilowatts, the steam consumption is 15.4 pounds per kilo- 
watt; and at 2,500 kilowatts, the steam consumption is 17 pounds per kilowatt. 
By this combination, the maximum output of the unit has been increased from 
1,265 kilowatts, non-condensing, to 2,500 kilowatts, or from 1,470 kilowatts 
condensing, to 2,500 kilowatts. 

It must not be thought that all this gain can be attained with 
no compensating loss. In the first place, a surface condenser, to 
maintain 28.5 in. vacuum, must be about twice the size of one to 
maintain 26 in., and requires special apparatus that is not only costly, 
but difficult to maintain. Again, the cost of maintaining a 28.5 in. 
vacuum is very much more than that of maintaining a 26-in. vacuum, 
leaving out of consideration the extra cost of condenser and cooling 
water. After all, it is the dollars and cents that determine the best 
efficiency, and it is poor economy to obtain the extra power at a 
greater cost than the returns will warrant. A gain of 35% or more in 
steam consumption may easily be effected by installing a low-pressure 
turbine, but the gain in dollars and cents is seldom as great; just what 
the gain may be, must of course depend upon the local conditions, 
especially upon the conditions under which the reciprocating engine 
is operating. In the majority of cases, such installations are worth 
while, even though used with the usual vacuum. 

An interesting application of the low-pressure turbine in con- 
nection with rolling mill machinery and other intermittent work, 
has been worked out by Professor Rateau, and has been made pos- 
sible by the use of his steam accumulator, or regenerator. This ap- 
paratus regulates the intermittent flow of steam exhausted from the 
rolling mill engine, let us say, and intended to be used by a low-pres- 
sure turbine. The accumulator may consist of a large tank in v/hich 
are numerous plates over which water can flow, or may contain simply 
water rapidly circulated by artificial means. As the exhaust steam 
from the engine enters this accumulator, it spreads out over the ex- 
posed water surface, and some of it is condensed if there is an excess 
of pressure due to more steam being supplied by the exhaust than is 
being utilized by the turbine. On the other hand, if the turbine 
utilizes more steam than is supplied by the exhaust, this causes a 
lowering of the pressure in the accumulator, and a rapid vaporiza- 
tion occurs from the exposed water surfaces, tending to equalize the 



40 



STEAM TURBINES 



pressure. The accumulator thus bears the same relation to the 
transfer of heat from the reciprocating engine to the turbine that a 
fly-wheel bears to the transfer of work from the cylinder of the engine 
to mill shafting. Fig. 27 shows one form of the Rateau accumulator. 
It must be provided, of course, with a safety-valve, set at a pre- 



Exhaust Steam from 
Winding Engine 



Oil Separator 




Basin in 
three pieces 



Bottom Basin 
in one piece 




Fig 27. Interior View of Rateau Accumulator, with Iron Trays. 

determined pressure, and is usually provided with a reducing valve from 
the boiler, so that in case the reciprocating engine should stop for a 
considerable length of time, steam could still be supplied to the 
turbine through the reducing valve. 

The first apparatus of this kind was installed in 1902, and has 
been very successful. The first to be installed in the United States was 
at the Wisconsin Steel Company, in South Chicago. In this plant, 



STEAM TURBINES 41 

steam first goes to a receiver to take out the shock due to the puffs 
of the exhaust. From here it passes to the regenerator. The 
receiver is fitted with baffle plates and drains for water and oil, by 
means of which they are thus separated from the steam. This 
accumulator at South Chicago furnishes steam for a low-pressure 
Rateau turbine which is used to furnish electric power for general 
purposes. 

Installation. The field of the steam turbine is unfortunately 
limited in its usefulness by two very important factors; first, its 
relatively high speed of revolution, even when compounded; and, 
second, its non-reversibility. If, as in marine work, reversing is 
absolutely necessary, then another turbine, which runs idle ordinarily, 
with vanes set in the opposite way must be fitted on the shaft. To 
make this reversing turbine as small as possible, efficiency is sacrificed, 
but this is of small consequence, for it is used so little. It of course 
adds materially to the first cost of the turbine and increases the 
length of the necessary floor space. 

The first and greatest field of turbine usefulness is undoubtedly 
central station work for the generation of electricity by direct-con- 
nected apparatus. It also has an important field in driving blowers, 
centrifugal pumps, etc., where high speed of revolution is essential. 
In such cases, it has a distinct advantage, for it may be direct-con- 
nected, thus doing away with the belting necessary if reciprocating 
engines were used. The turbine has been suggested to some extent 
for driving mill shafting, in which case, of course, the speed is belted 
down from a small pulley on the turbine to a large one on the counter- 
shaft, but this appears to offer no particular disadvantage, for in 
any case belting would be used, as the countershaft would never be 
run at the same speed as the ordinary reciprocating engine. 

In the field of electric generation the turbine to-day has prac- 
tically superseded the reciprocator. The number of installations 
is very great, and probably no new central station is now designed 
for other than steam turbines. In 1906 the Committee of the Na- 
tional Electric Light Association, after an extensive investigation 
of turbines, reported a wide use of turbines for electric generation, 
and their figures showed that about 75% of all the turbine units 
of 500 kw. or over already installed in the United States, were for 
electric purposes, and that practically only one new central station 



42 



STEAM TURBINES 



abroad had been found installing reciprocating engines. The 
distinct advantage of turbines for this work is the uniform turning 
effort, the high speed of rotation permitting the use of a very much 
smaller generator, and the smaller floor space, requiring less capital 
outlay in land and engine-house. These features place it in striking 
contrast with the ponderous slow-moving Corliss engine. 

The General Manager of the Metropolitan Street Railway Co. 
of Kansas City is authority for the statement that in that station 
six 5,000 kw. units of a well-known make of turbine could be installed 
space previously occupied by three 3,000 kw. engine-driven 



in 



Fig. 28. 



e & e q 

e e o e 




Plan and Elevation of 500-K. W. Westinghouse Turbo-Generator. 
Same Scale as Fig 29; Notice Difference in Space Required. 



This is 



units. Or in other words, 30,000 kw. of turbine power could be 
put into a building where before only 9,000 kw. of engine power 
had been possible. This probably is greater than would ordi- 
narily be met with, but the difference in any case is large, the sav- 
ing in space depending upon the type of turbine. The average 
horizontal turbine and generator with auxiliary apparatus will occupy 
about three-fifths of the space needed for a slow-speed, engine-driven 
generator of the same power, and a vertical turbo-generator somewhat 
less space than the horizontal. 

A further distinct advantage of the turbine is in the fact that, 
since there are no valves to adjust, the efficiency can be lowered only 
by wear, and then only slightly; on the other hand, in reciprocating 
engines, if the valves are not set exactly right, very poor economy 



STEAM TURBINES 



43 



will result, and the opportunities for wear are far greater than in 
turbine engines. Again, the turbine can use high degrees of super- 
heat because there is no lubricant to burn; there is also little danger 
of entrained moisture in the steam wrecking the turbine, and the 




O S f 6 6 tO 
•Scale or feet 





f @ © ©* 


i- 1 -J u 


; .• i 


U J ° 


O 1 5 ^ +2l 




Fig. 29. Plan and Elevation of 500-K. W. Corliss Engine-Driven Generator Set. 
Compare with Fig. 28. 



absence of oil in the condensed steam greatly lessens trouble in the 
boiler if the condensation is used for feed water. The economy of 
space was graphically illustrated by Fig. 1, and Figs. 28 and 29 tell 
the same story but with different types of engine and turbine. 



44 STEAM TURBINES 

Figures showing the relative space occupied by reciprocators 
and turbines are of little value unless the size of condenser and 
condensing auxiliaries are taken into consideration, for, as before 
mentioned, they may easily be, in the case of the turbine, twice the 
size of those used with a reciprocating engine of the same power. 
The apparent saving of space, therefore, may be offset by these 
auxiliaries. By placing the condensers underneath the turbine, 
as is frequently done at the present time, not only may a consider- 
able amount of floor space be saved, but the turbine can more readily 
exhaust into the condenser. As we have already seen, at high vacuum 
the volume of steam is very large, and the exhaust pipe from the 
turbine will be proportionally large. It would thus appear that to 
have the condenser any great distance from the low-pressure end of 
the turbine would be not only a distinct disadvantage, but offer a 
considerable practical difficulty. 

Turbines, as we have seen, require very much smaller founda- 
tions than reciprocating engines of the same power, and these founda- 
tions will therefore cost very much less. It is hard to get a direct 
comparison between turbines and reciprocating engines as a class, 
because the foundations for high-speed reciprocating engines will 
not be as massive as for the heavier, low-speed engines. The tur- 
bine, occupying less floor space, will require smaller buildings and 
less land, and this will in a number of cases be a substantial saving 
in first cost and subsequent interest charges. 

So far as the first cost of a generating plant goes, there is at 
the present time very little difference between those using reciproca- 
ting and those using turbine engines. The turbine itself costs more 
than the reciprocating engine of the same power, but on the other 
hand the generator for the turbine costs very much less. Again, 
the condenser and pumps, if high vacuum is to be maintained, will 
cost two or three times as much as for the reciprocating set, while 
the cost of erection is decidedly in favor of the turbine. It is not 
easy to get a direct line on the relative cost of turbine and engine 
installations, for the figures available appear to vary about as much 
between reciprocating engines and turbines as might be expected 
to be found between various installations of reciprocating engines, 
and undoubtedly turbine installations in some cases cost relatively 
more than in others. It seems probable that the cost of the turbine is 



STEAM TURBINES 45 

regulated more by the cost of the reciprocating engine with which 
it has to compete, than by the actual cost of manufacturing the 
turbine. All in all, there is likely to be a somewhat less cost of com- 
plete installation in favor of the turbine, but the difference will not 
be large in any case, and in powers under about 100 kw., it is 
probable that the engine installation is fully as cheap. This does not 
take into account the value of land and buildings, which in all cases is 
an important factor in favor of the turbine. 

Performance. The losses occurring in the steam turbine con- 
sist principally of loss of velocity of the steam itself due to friction 
in contact with the vanes and guides; friction of the disks revolving 
through a chamber filled with steam; eddying of the steam jet, due 
to improper speed of the revolving disks; radiation; bearing friction. 
The two latter items are not large, and under ordinary conditions 
would consume less than 2% of the power. 

The most important losses are due, first, to the friction of the steam 
jet against the vanes and guides, which will be approximately propor- 
tional to the cube of the velocity of the steam relative to the vanes 
or guides, and second, to the considerable amount of friction of the 
disks as they revolve in the chamber filled with steam. This friction 
generates heat which raises the temperature of the steam and metal 
parts and thus causes the re-evaporation of some of the condensed 
moisture. Since this adds some heat to the expanding steam, the 
expansion is not absolutely adiabatic. The smoother the revolving 
wheels are made, the less will be this friction, a fact well illustrated by 
a reported improvement of about 1% in steam consumption which 
was effected in a well-known make of turbine by making the riveting 
on the revolving disks perfectly flush. To these losses may properly 
be added the generator losses which, of course, are a factor of the 
speed of revolution. 

With either reciprocating engines or turbines, the steam economy 
is much better in large than in small units, and especially is this 
true of the turbines of the reaction type. In small turbines of this 
type, the steam friction is high and the leakage large, and this makes 
it undesirable to build this type of turbine in sizes much below 500 
kw. For the impulse type of turbine, these losses are not as impor- 
tant in the smaller powers, and DeLaval, Curtis, and Rateau turbines 
of comparatively small power can be built to give nearly as good 



46 STEAM TURBINES 

steam economy as larger turbines of the same type, and can easily 
excel small reciprocating engines. 

The steam consumption of the turbine depends naturally 
enough upon the vacuum, steam pressure, degree of superheat, 
variation in load, and variation in speed. It has already been 
explained that the turbine can utilize the lower ranges of vacuum 
far better than can the reciprocating engine, but it could not, in all prob- 
ability, use the higher pressure ranges with as good economy as the 
best reciprocating engines. If the turbine runs at a vacuum of 27 in., 
its steam consumption will be practically on a par with that of the 
reciprocating engine, and it will show a gain of about one-half pound 
of steam per kw.-hr. for each extra inch of vacuum, below 25 in. But 
from the saving effected by this one-half pound of steam must be 
deducted the extra cost of maintaining the high vacuum, if the real 
economy is desired. Not only can the turbine theoretically utilize 
the greater vacuum to better advantage, but it has an advantage also 
in a practical way, because with the reciprocating engine, a very high 
vacuum cools the cylinder walls and thus causes a relatively large 
initial condensation, which difficulty is not met with in the turbine, 
the high vacuum having no detrimental effect. It thus has both a 
theoretical and practical advantage. 

Superheated steam, whether used in the reciprocating engine 
or in the turbine, will reduce the steam consumption; but in the re- 
ciprocating engine, superheating cannot be carried very high, as 
the cylinder lubricant is likely to be burned, and there will be little 
condensation in the cylinder to help out the lubrication. The tur- 
bine is not handicapped in this way, but nevertheless high degrees 
of superheat are likely to cause trouble due to unequal expansion in 
the casing, the temperature at the high-pressure end being so much 
greater than that at the low-pressure end. This expansion is trouble- 
some, but should be provided for in the design. 

Superheat affects the economy of the steam engine in two ways ; 
it carries additional heat units into the cylinder, and lessens con- 
densation. It also helps in the turbine in two ways; it carries 
additional heat into the turbine, and, being less dense than saturated 
or moist steam, causes less friction within the turbine, and thus 
effects a mechanical as well as a theoretical gain. It is generally 
reported that the gain is 10% for each 100° of superheat, but tests 



STEAM TURBINES 47 

which appear to be thoroughly reliable do not seem to bear out this 
claim. 7J% to 8% is a better figure. 

The saving in steam will be from 1.5 to 1.75 pounds per 
kw.-hr. for each 100° of superheat, but the real economy resulting 
from this superheat will be the difference be ween the value of 
this saving in steam and the cost of superheating. The superheating 
plant costs more, not only for the additional expense of the super- 
heater, but for piping, valves, etc. Cast-steel fittings, and valves 
with nickel-steel valve stems, are usually required for high degrees 
of superheat. 

The usual steam pressure in turbine work is about 150 pounds 
gauge. If lower than this, some gain in steam consumption may be 
had by an increase in boiler pressure, but an increase over 150 pounds 
does not appear to be productive of great economy. A reference to 
Fig. 25 will readily show that increasing the pressure above 150 
pounds will add very little to the area of available work. Fig. 30 
shows the curves of economy of a 30-H. P. turbine at different steam 
pressures. The gain is less and less the higher the pressure becomes, 
and is small from 75 to 100 pounds. From 35 to 100 pounds the 
gain is about 33 J%, but this gain is not due entirely to the rise in 
steam pressure. 

The study of steam nozzles has shown that to use steam efficiently, 
the nozzle must be properly designed v/ith reference to both the 
initial and final pressures. Now, if the nozzle on this turbine were 
designed for 100 pounds pressure, it could neither utilize steam 
economically at 35 pounds, nor at 150 pounds pressure. To show 
the real gain due to an increase in steam pressure, it would be 
necessary to have nozzles in each case that were designed for the 
specific pressures used. Then, and only then, would the curves 
show the true gain due to increase in pressure. But a study of 
Fig. 25 shows that if the theoretical gain is small the practical gain 
cannot be large. It must, moreover, be borne in mind that a high- 
pressure plant costs more than a low-pressure plant, and for 
stationary work very high pressures will not pay. On shipboard, 
where space and weight are at a premium, it may be good engineering 
policy to install very high pressures, even though the first cost is 
greater. 

Fig. 31 shows the curves for a 600-kw. Curtis turbine with vary- 



48 



STEAM TURBINES 



ing pressures. In this type of turbine, the same conditions exist as 
in the previous one, the nozzles being designed for only one pressure. 
The economy of a turbine varies with the load, as does the econ- 
omy of the reciprocating engine, but not perhaps to as marked an 
extent, and the economy depends of course upon the type of tur- 
bine. Turbines like the DeLaval and Curtis admit steam through 
a number of nozzles which are opened and closed either automatically 
by the governor or by hand. At normal load, about two-thirds of these 
nozzles would be open and a 50% overload could then be carried 
with all nozzles open. In the Parsons turbine, steam is admitted 
all around the circumference of the drum but the admission is in- 
termittent. For heavy loads the valve remains open for longer in- 



so 



20 



* 








3 


H. 


P. S 


tear, 


i TO 


rbin 


e Mo 


tor 








35 

0) 






























si 

«0 












































v5 

1 — °~1 


15 lbs 

I 1 


. ste< 

l 


im pi 

l 


■essu 

p — o 


re 


















i 50 /As. 








75lbs. » 


i-- - *? 
















/OO/bs 
1 






t=o- 






■3 

-J 












B 


rake 

L. 


1 
' H.P. 

1 













30 
Fig. 30. Curves Showing Economy of 30-H. P. Turbine. 



35 



tervals, and when the load is such that the valve remains open all the 
time, further overloads can be provided for only by resorting to a by- 
pass which admits high-pressure steam to the second stage of the 
turbine. In such cases, of course the economy falls off, for the 
steam does not get the benefit of full expansion. At low loads, there 
is not a great deal of choice between the different types of turbine, 
but those that can carry a large overload without opening a by-pass 
are bound to be the most economical under these conditions. 

Overload is taken care of in a reciprocating engine by increasing 
the cut-off, but, as this reduces the number of expansions, this method 
is uneconomical. For small ranges of load, the relative economy of 
turbine and reciprocator are not very different, but the effective 
range of the turbine is much greater than for the reciprocating engine. 



STEAM TURBINES 



49 



A good turbine will carry 100% overload for a short time and will 
carry 50% to 60% overload on approximately 10% more steam. 
Fig. 32 shows characteristic curves of steam consumption at varying 
loads. 

A variation in speed of the turbine within moderate limits does 
not materially affect the economy. The best speed of the vanes 
(see Page 17) is half the velocity of whirl (h V cos a). Moderate 
departures from this speed do not materially affect the economy, 
provided the entrance angles of the vanes are such that the steam 
jet strikes without shock. The angle of the vane must depend upon 
the speed, and once fixed, 
any variation in speed 
must of course cause the 
steam jet to spatter and 
form eddies, a source of 
material loss. This is 
entirely apart from the 
question of whether or not 
the designed speed of ro- 
tation is the most econom- 
ical. To avoid spatter- 
ing and eddy losses, the 
vane angle must change 
with the speed, which is 
evidently impossible 

A rapid change in load will cause cylinder condensation in a 
reciprocating engine, so that, on test under steady load, the engine 
is likely to show up better than it would under service conditions. 
With the turbine, this is not so. Here there is no such condensa- 
tion, and the performance under test is far more likely to agree with 
performance under service conditions. Both types of motor will 
fall off under service conditions, but if an engine and turbine do 
equally well under test, under such widely varying conditions as 
exist in a central station, for instance, the turbine ought to show 
up better in actual service. 

A reciprocating engine is usually designed for a low average 
load and, therefore, it will permit a relatively large increase in load, 
but it is generally working on a slight underload, and hence at less 



21 




\ 












20 




\ 












S 


\ 














* 














/9 
































<0 


















/n 


tiaf pr 


essure. 


lbs. per 


Squort 


? inch. 



Fh 



/20 MO /60 J80 200 220 

31. Economy Curves of 600-K W. Curtis Tur- 
bine with Different Steam Pressures. 



50 



STEAM TURBINES 



than the maximum efficiency. The turbine, on the other hand, is 
usually designed for its normal and most economical load, taking 
care of overload by opening more nozzles at theoretically the same 
efficiency, or by opening a by-pass at somewhat less efficiency. This 
should give the turbine a still further advantage at the end of the 
day's work. 

Tests. Tests of reciprocating engines usually give steam in 
pounds per indicated horse-power per hour, but there being no 



! 



I* 

r 













































i 


7 

f 


/ 
































/ 




























« 


il 


/ 






























1 


K 


/ 






























< 


i 


1 






























< 


i 


1 


/n 


pulse Typ 


? n 


irbine 


















\ 


X 


| 




















— 


— - 


-■ 


._, 


1 


\ 


Norma 


/ loads 


approxim ateli/ 






at int 


er section 


of curyes 
















H \ L 
































i 


































\ \ 


\ 
































\ 


\ 


K 






























\ 


V 





































\ 




































s 


s 

















































o 



ts 



13 



/4 



/5 



/6 



/7 



/a 



/9 



20 



Fig. 32. Economy Curves of Turbines and Compound Corliss Engine with 
Varying Loads. 



indicated horse-power for a turbine, the comparison must be made 
on some other basis. Brake, or shaft horse-power may readily be 
obtained for a turbine, and in engine tests, where the brake horse- 
power has been determined, there is of course opportunity for a 
direct comparison. However, since engineers are in general more 
familiar with steam rates per I. H. P., it seems well to consider 
how a comparative I. H. P. may be had for the turbine. Various 
tests to determine the relation between brake and indicated power 



STEAM TURBINES 51 

on reciprocating engines seem to show that 92% is a fair figure 
for a good engine. 92% then of the steam rate per brake horse- 
power would give the rate per comparative indicated horse-power. 

The largest field for the steam turbine being central station 
work, it follows that by far the larger number of turbine tests are 
quoted in terms of electrical units. It is costly to fit a brake for 
a large turbine and entirely useless when the power delivered at the 
switchboard can be read off at once. For electrical work, of course 
reciprocating engine tests are often quoted in the same electrical 
units, in which case, there are abundant opportunities for direct 
comparison. 

Suppose, however, that it is desired to compare steam per 
I. H. P. with a corresponding rate per kilowatt-hour at the switch- 
board. 1 kw. = 1.34 electric horse-power measured on the switch- 
board, which is evidently shaft or brake output less losses in the 
generator. Since the efficiency of a good generator is not far from 
95%, the brake horse-power will be equal to the electric horse-power 

95 
divided by -rp^r* We may say, therefore, in ordinary cases, that 

1.34 X kw. 
B ' R R = ^5~ 

, B. H. P. 92 
since we assume, that ^ ^ = ;— r. 

we have I. H. P. = * n * = 1.53 kw. approximately. 

Steam per kw.-hr. then, divided by 1.53 would give the steam per 

comparative iridic ated horse-power per hour, or 

steam per I. H. P.-hr. Xl.53 = steam per kw.-hr. 

A turbine using 20 lbs. of steam per kw.-hr. would be about on a par 

20 
with a reciprocating engine using — ^ =13 lbs. per I. H. P. 

In comparing the performance of one engine with the perform- 
ance of another, or one turbine with another, or an engine with a 
turbine, pounds of steam per horse-power per hour is generally the 
rough basis of comparison, but this is very crude and often mislead- 
ing. For instance, one test may be made with superheated steam 
and another with saturated or even moist steam, or one may have a 
higher steam pressure, or the vacuums may be different. 



52 



STEAM TURBINES 



To get an approximately intelligent comparison, all tests should 
be reduced to a standard degree of superheat, pressure, and vacuum, 
or better still, if the comparison is between two, correct both to the 
average conditions of the two. The corrections applied are more 
or less arbitrary, and it is manifestly unfair to apply them all to 
either test. If each is corrected for half the difference, a much more 
reliable comparison is likely to result. 

It is generally accepted that the steam consumption will de- 
crease about 8% for each 100° of superheat, about 5% for each inch 
of vacuum below 28 in., and about 5% for 50 lbs. rise in steam pres- 
sure between 100 and 150 lbs., and 3% for similar rise between 
150 and 200 lbs. The manufacturer usually gives guarantees of 
steam rates for various pressures, vacuums, and degrees of super- 
heat. When such figures are available, it is probable that their 
use would lead to more satisfactory results than if the rough approxi- 
mations mentioned above were used, but such figures would be cor- 
rect only for the one individual turbine, and in the large majority 
of cases the engineer is compelled to use the approximations. They 
are in most cases fair and satisfactory in the absence of definite data. 

To illustrate this method, consider a turbine at 177.5 lbs. (gauge) 
steam pressure, vacuum 27.3 in., superheat 96° F., consuming 15.15 
lbs. steam per kw.-hr., and another using 179 lbs. steam pressure, 
29.5 in. vacuum, and 116°F. superheat, consuming 13 lbs. of steam 
per kw.-hr. The average conditions are 178.2 lbs. steam pressure, 
28.40 in. vacuum, and 100° F. of superheat. 

The work will appear clearer if arranged in tabular form as 
in Table I. 



TABLE I 
Steam Consumption Tests 





TURBINE^l 


Turbine ^2 


Average 
Conditions 


Turbine 41 
Correction 

/o 


Turbine 42 
Correction 

% 


Steam bt Gauge 

Vacuum 

Super-heat 


177.5 lbs. 
27.3 in. 
96°F. 

15.15 lbs. 
14.19 lbs. 


179 lbs. 
29.5 in. 
116°F. 


178.2 lbs. 
28.4 in. 
106°F. 



-5.5% 
-0.8% 




+ 5.5% 
+ 0.8% 


Steam ) Observed 
per > 
Kw.-Hr. ) Correct'd 


13 lbs. 
13.82 lbs. 




-6.3% or 

-0.96 lb. 


+ 6.3% or 
+ 0.82 lb. 



STEAM TURBINES 53 

The correction for steam pressure, being only for .7 lbs., is too 
small to be of consequence in this case. The vacuum correction is 
±1.1 inches, and at 5 % per inch (the decrease in steam consumption 
for each inch of vacuum, as explained on Page 52), the correction would 
be ±5J%. The superheat correction is for 10°, or, as the decrease 
in steam consumption for 100° of superheat is 8%, this will be t-'tt of 
8% = 0.8%. The sum of these corrections gives ±6.3%, making 
.96 lbs. to be subtracted from turbine #1, and .82 lbs. to be added to 
turbine #2. The final steam consumptions, then, which should be 
compared are 14.19 lbs. and 13.82 lbs. instead of 15.15 lbs. and 13 
lbs. Turbine #2 appears, therefore, to use about 3% less steam 
than turbine #1 under similar conditions. 

Another and perhaps more satisfactory method of comparison 
is by means of the heat units used. This computation may be made 
readily from the steam tables. Using the same tests as given above, 
turbine #1 uses steam at 177.5 lbs. gauge pressure = 192.2 absolute, 
at which pressure each pound of dry saturated steam contains 1197 
B. T. U. If we allow | B. T. U. for each degree of superheat, then, 
for 96° F. we should add 96 X. 5 = 48 B. T. U., and each pound 
would then contain 1245 B. T. U. at admission. If this steam is 
condensed at a pressure of 27.3 in. vacuum = 1.33 lbs. absolute, 
each pound of the condensation will contain 80 B. T. U. which will 
be returned to the boiler in the feed water. The net amount then 
consumed by the turbine and carried away by the cooling water 
of the condenser is 1245 - 80 = 1165 B. T. U. per pound. 15.15 
lbs. would represent 15.15 X 1165 = 17,650 B. T. U. per hr. or 294 
B. T. U. per kilowatt per minute. 

Turbine §2 uses 13 lbs. of steam at 179 lbs. gauge pressure and 
116° F. superheat, condensing at 29.5 in. vacuum. In this case, each 
pound of dry steam at admission would contain 1197.3 B. T. U. 
and 116° F. superheat would add about 58 B. T. U. more, making 
1255.3 B. T. U. per pound. Condensing at 29.5 in vac. = .25 lbs. 
absolute, each pound of condensed water would contain 27 B. T. U. 
to return to the boiler in feed water, leaving 1255.3 — 27 = 1228.3 
B. T. U. to be used by the turbine. 13 lbs. would represent 

— — — = 266 B. T. U. per min. to compare with 294 in the 

previous case. 



54 STEAM TURBINES 

Here, again, the direct comparison is likely to be misleading, 
unless due account is taken of the difference in conditions. The 
gain is apparently about 10% in favor of turbine #2 on the heat unit 
basis taken under the actual working conditions of each, but the fact 
must not be lost sight of, that turbine #2 is working under more favor- 
able conditions of vacuum and ought to show a much better efficiency. 
It appears from this discussion that both turbines work under the 
conditions of design with but little difference in actual economy. 

Turbine manufacturers are in the habit of reporting tests of 
the turbine only, no account being made of the auxiliary apparatus. 
This is manifestly misleading, for with a 29-in. vacuum, the power 
consumed by auxiliaries may easily be twice what it would be for a 
27-in. vacuum. This extra power and the cost of maintaining it 
in a measure goes to offset the gain due to the higher vacuum. 




o 

O o 

ft J 



° I 
a § 
o t> 
o 
w 



STEAM TURBINES 

PART II 



COMMERCIAL TURBINES 

*In this description of commercial turbines it will be convenient 
to classify them as follows: 

Single-Stage Type 

(Compounding by Velocity Steps only 
I Compounding by Pressure Stages only 



I. Impulse 
Turbines' 



Compound Type 

| Compounding by both Velocity Steps 
[ and Pressure Stages 

II. Reaction Turbine 

III. Combined Impulse and Reaction Turbine 



IMPULSE TURBINES 

SINGLE=STAGE IMPULSE TURBINES 

Probably the simplest type of turbine is the one with a single 
stage, that is, a single set of nozzles and a single rotating wheel, 
but, as already pointed out, the velocity of rotation in a turbine of 
this sort is usually so great that some device must be employed to 
reduce the rotational velocity. This may be done in two ways. 

As has been previously stated, the feature of importance is 
not the rotative speed but the peripheral velocity of the wheel, which 
is somewhat less than one-half the steam velocity. Maintaining 
this peripheral velocity constant, turbine rotors of comparatively 
small diameter may be used, the high rotative velocity being reduced 
by means of gearing; or, the diameter of the turbine rotor may be increased, 
the rotative speed thereby being reduced in the same ratio that the 
diameter of the wheel is increased. 



*Many writers group by themselves all turbines using buckets of the Pelton type, but 
this does not seem to be a proper classification, as it is the action of steam in the turbine that 
makes it belong to a certain type, and not the style of bucket that is used. Turbines using Pelton 
buckets may belong to any of the impulse groups. 



56 STEAM TURBINES 

Both of these methods have been employed in turbines which 
have been put on the market, the first method being characteristic 
of the De Laval turbine, and the second, of the earlier forms of the 
Riedler-Stumpf machine. The manufacturers of the latter dis- 
carded this scheme in their later designs in favor of a compound 
turbine of some sort. 

De Laval Turbines. The turbine designed and developed by 
Dr. Gustav De Laval of Sweden was among the first to be commer- 
cially successful. His first turbine, which was used to run the famous 
De Laval cream separators, was of the pure reaction type, similar 




Fig. 33. Principle of Operation of De Laval Steam Turbine 

in action to the old Hero engine. This turbine was not economical 
in steam consumption, but, as it was used for very small powers only, 
this factor was not important, and commercially, the machine was 
very successful. This success led to the desire to build larger tur- 
bines, and in developing them the reaction principle was abandoned. 
The essentials of the motor element of the De Laval turbine are 
Illustrated by their familiar trade-mark, shown in Fig. 33. They 
consist of a rotating disk, having vanes on its periphery; a number of 
hozzles in which the steam is expanded from boiler pressure to the 
pressure in the exhaust chamber and delivered in a jet against the 



STEAM TURBINES 57 

vanes; a shaft to which the rotating wheel is fixed, so arranged that 
at high speed the rotating element can revolve about its own center 
of gravity* instead of its geometrical center; and a set of reducing 
gears to reduce the high rotative speeds to the desired amount. 
It is an impulse turbine with a single wheel carrying one row of 
buckets, and is a single-stage turbine in all respects. The steam is 
directed against the vanes from nozzles with flaring sides, which 
are so designed as to give it the maximum velocity and to expand it 
within the confines of the nozzle to the pressure in the exhaust 
chamber, thus transforming all of the heat energy of the steam into 
kinetic energy. The nozzles deliver the steam jets at the side of the 
wheel, and for a maximum efficiency should make as small an angle 
as possible with the plane of rotation.! In the De Laval machine 
this angle is 20 degrees. 

For small turbines, the entrance and exit angles of the vanes 
are 32 degrees, increasing to 36 degrees for the larger sizes. Under 
these conditions the best peripheral velocity will be about 1900 feet 
per second while the velocity of the steam issuing from the nozzles is 
4000 feet per second. In most impulse turbines the peripheral 
velocity varies from 1400 in the larger sizes to 500 in the smaller 
sizes. These speeds are high, even for turbine work, and necessitate 
the solution of very interesting engineering problems. These 
velocities, with the diameters used for De Laval machines, mean 
revolutions of about 10,600 per minute in the larger sizes and 30,000 
per minute in the smaller sizes, these speeds being reduced by helical 
gears to approximately 900 and 3000 revolutions per minute, respec- 
tively. 

In the small-sized turbines this gearing consists of a pinion and 
a single gear, but in the larger-sized turbines there is a single pinion 
with a gear on each side. This method has the advantage of dis- 
tributing half the load on each gear, thus lowering the pressure on the 
teeth and eliminating the side pressure on the bearings of the flexible 
shaft. 

Nozzles. The nozzles are set in the casing in which the wheel is 
enclosed and are opened and closed by means of hand valves. A 

* Lack of uniformity in the density of the steel might cause the center of gravity of the 
wheel to be outside of its geometrical axis. 
tPage 17, Part I. 



58 



STEAM TURBINES 



detail of the nozzle and valve is shown in Fig. 34. A is an annular 
space in the casing acting as a steam chest, C is the valve which 
permits opening or closing of the nozzle, and B is the nozzle itself. 
The nozzle is fitted into a taper hole in the casing and drawn into 
place by a nut. 

The design of the nozzle naturally depends upon the pressure 
used, the degree of superheat, and the vacuum or back pressure. 
The nozzles being easily removed, it is apparent that a turbine can 
readily be altered to meet different conditions by inserting new 
nozzles. A condensing turbine is often equipped with an extra set 
of nozzles designed for non-condensing conditions, which may be 
used with better economy in case the vacuum fails. 




Fig. 34. Detail of Nozzle of De Laval Turbine 



There are usually from 2 to 24 nozzles in the casing, and the 
power developed at any time naturally is proportional to the num- 
ber of nozzles in operation. The clearance between the wheel and the 
nozzle is about f of an inch. The clearance Ibetween the tips of the 
blades and the casing is not a matter of importance, for there is no 
tendency to steam leakage, the pressure in all parts of the casing 
being practically the same as the back pressure. This clearance, 
therefore, may be whatever practical conditions require. Fig. 35 
is the exterior view of the turbine and generator, showing nozzles 
and valves set in the casing. By inserting nozzles in the holes which 
are shown plugged in the figure, a greater power can be obtained. 

Vanes. The vanes are of the crescent shape common in impulse 
turbines. They are made of drop-forged steel which resists erosion 



STEAM TURBINES 



59 




g C 



Q-S 



o 9 



5.3 



o a 



G°s 



60 



STEAM TURBINES 



and have bulb shanks, as shown in Fig. 36, which are driven into 
place. The outer ends of the vanes fit close together, thus form- 
ing a continuous ring, which prevents any movement at the ends 
of the vanes. 

Steam at high velocities, especially if wet, is liable to cause 
appreciable wear on the vanes, the wear being practically all on the 
entrance side; but it is not very great, and tests of a 100-horsepower 
turbine have shown that wear on the buckets could be as great as 
fg of an inch without increasing the steam consumption more than 
3 per cent, according to the report of the manufacturers. 

Wheel. At the very high speed employed, centrifugal forces 
are enormous, hence, special high-grade nickel steel must be used in 

the manufacture of the 
rotating elements. The 
steel is said to be high 
in carbon and to pos- 
sess a tensile strength 
of approximately 135,- 
000 pounds per square 
inch. The wheel is 
shown in cross section in 
Fig. 37 and is designed 
to be of uniform strength 
throughout, except just 
below the rim, where a 
narrow annular groove is turned purposely to make this section weak, 
for the following reason : 

Centrifugal force increases as the square of the speed, and should 
the safety devices fail to work, the rotating wheel must ultimately 
burst. The reduced section near the periphery of this wheel makes 
the stresses at this point approximately 50 per cent greater than 
elsewhere, and yet, at normal speeds, this will be perfectly safe, as the 
factor of safety is between 4 and 5. Now, since the centrifugal 
force increases directly as the square of the number of revolutions, 
the stresses at the weakened point, when the speed is double, will 
be four times as great, that is, about equal to the ultimate strength 
of the material. Consequently the rim will burst and fly into many 
small pieces, doing but little damage, however, as the casing is made 




Fig. 36. Drop-Forged Vanes and Method of Attach- 
ment in De Laval Turbine 



STEAM TURBINES 



61 



heavy enough to restrain these fragments. When the rim flies off, 
the stresses in the main portion of the wheel are thereby greatly 
reduced and no further damage can ensue. Wheels without this weak 
section have burst under experimental tests into a few large pieces pos- 
sessing enough energy to break through a 2-inch cast-iron casing. 
On each side of the wheel are hubs extending into cylindrical 
openings in the casing. These are known as safety bearings and 
work with slight clearance under ordinary conditions. Should the 
rim burst, the wheel would at once become unbalanced and the result- 




b- — my— 





Fig. 37. Method of Mounting Wheel of Small De Laval Turbine 



ing eccentricity of the center of gravity would cause the wheel and 
shaft to rotate off-center, bringing a considerable pressure of the hub 
against these safety bearings. These acting as a brake, together 
with the absence of further impelling forces due to the loss of rim and 
buckets, will quickly bring the rotating wheel to a stop. 

For small wheels a bushing is fitted and shrunk to a short swelling 
on the shaft and, in addition, is pinned in place. The hub of the 
wheel is bored to fit this bushing and, together with the shaft, is 
drawn into place by a nu+, as shown in Fig. 37. The wheel may 
readily be removed from the shaft by loosening the nut. 



62 



STEAM TURBINES 



For large wheels such a construction is not desirable, because a 
wheel with a hole in the center is not nearly so strong as one without 
such a hole, and in the larger sizes of turbine the strength of the 
wheel is an exceedingly important factor. The hub, therefore, in such 
a wheel is solid, but is recessed to fit the flanged end of the shaft, as 
shown in Fig. 38. The recess is tapered \ inch to the foot, to fit the 
shaft, which is securely bolted in place, as shown. The rim of the 
wheel is drilled parallel to the shaft, with cylindrical holes milled out, 




Fig. 38. Method of Mounting Wheel of Large De Laval Turbine 

as shown in Fig. 36, to hold the bulb shanks. This makes a strong 
construction, and the vanes are easily replaced when necessary. 

Shaft. When a body is rotating at high speed, it must be very 
carefully balanced by distributing the material symmetrically about 
the center of rotation. If the center of gravity of the rotating mass 
is not absolutely at the center of the shaft, a vibration more or less 
serious will be set up, because a rotating body tends to rotate about 
its own center of gravity instead of its geometrical center, thus caus- 



STEAM TURBINES 



63 



ing a pressure alternately on one side or the other of the bearing. 
For speeds of 3000 r.p.m., which are common in compound turbines, 
the wheels can be balanced on knife-edges, the wheel disks being 
drilled at certain points until they become perfectly balanced. It 
is claimed that careful work in this matter will ensure the center of 




gravity of the wheel being within T <rV o of an inch of the geometrical 
center. Small as this error may be, it would be prohibitive at the 
high rotative speeds used in the De Laval turbine; hence the adoption 
of the long, slender shaft on which the wheel is mounted. This bends 
slightly, and allows the wheel to rotate about its own center of 



64 



STEAM TURBINES 



gravity without vibration. This feature is distinctive of the De Laval 
machine. The relatively small diameter of shaft is astonishing, 
being only a little over 1J inches at its smallest section for the 300- 
horsepower turbine. 

Gears. The speed-reducing gears are in the ratio of about 10 
to 1; i.e., if the turbine rotor has a speed of 30,000 r.p.m., the 
larger gears have a speed of 3000 r.p.m. At the desired place, a 
swelling on the shaft is provided in which the pinion teeth are cut. 
In the smaller sizes only one large gear is used, but in the larger 




Fig. 40. Horizontal Section of De Laval Single-Geared Turbine 
A, Wheel Case; B, Wheel Case Cover; C, Turbine Wheel; D, High-Speed or Pinion Shaft; F , 
Outboard Bearing Bracket; I, Outboard Ball-Seated Bearing; J, Outer Packing Bushing; K, 
Inner Packing Bushing; L, Gear Case; M, Gear; O, Gear Shaft Bearings; P, Inner Pinion Bear- 
ing; Q, Outer Pinion Bearing; R, Vacuum Governor Air Valve; S, Governor; T, Nozzle Chamber: 
V, Exhaust Chamber; W, Gear Shaft. 

machines there are two large gears, one on each side of the pinion. 
The teeth are cut spirally at an angle of about 45 degrees, as shown 
in Fig. 39, and have double sets of teeth at 90 degrees to each other. 
These reduction gears are fine examples of engineering and 
mechanical skill, as only the best work would stand up under the 
high speeds of rotation. The shaft on which the pinion is cut is of 
nickel steel, but the gears are made of soft steel, low in carbon. They 
have a peripheral velocity of about 100 feet a second and, if kept 
free from grit, will run for a long time with little or no wear. These 
gears were originally made of bronze, but this metal proved unsatis- 



STEAM TURBINES 



65 



factory because of the crystallization which it developed, and which 
resulted in the fracture of the teeth after a few years' continuous use. 

Fig. 39 illustrates the various working parts of the De Laval 
steam turbine. B is the rotating bladed wheel, A the long flexible 
shaft, C the pinion cut on the shaft, H one set of reducing gears, 
and m the flange for connection to the working unit. Fig. 40 
shows a sectional view of a complete turbine and connections for a 
single working shaft. 

Bearings. The turbine shaft is supported in three bearings. 
The outer bearing is solid and is held against a ground spherical 




Fig. 41. Details of De Laval Governor and 
Automatic Safety Vacuum Valve 



Fig. 42. Details of Bell-Crank Lever and 

Throttle Valves for Governor Shown 

in Figure 41 



seat by means of a cap and spring. It is made of bronze, lined with 
babbitt metal. The other two bearings are arranged one on each 
side of the pinion. They are very long, and hold the pinion accu- 
rately in mesh with the gear. They are split to facilitate removal, 
and have suitable grooves for lubrication. No provision for adjust- 
ment is made. The lubrication is supplied from a central reservoir 
by means of sight-feed lubricators. 

Governor. The speed regulation is obtained by means of a simple 
type of centrifugal governor located, in the geared type of turbine, 
at the end of one of the slow-speed shafts. It consists of two weights 
D f Fig. 41, hinged on knife-edges, acting on a sliding collar J, mounted 



66 



STEAM TURBINES 



on a spindle 7. The governor weights in moving outward push against 
the collar, moving the spindle outward and at the same time com- 
pressing the springs H. The spindle in turn presses against the 
pin N at the end of bell-crank lever L, Fig. 42. This bell-crank lever 
operates the throttle valves G and F shown in Fig. 42. 

Riedler=Stumpf Turbine. The first turbine developed by Pro- 
fessors Riedler and Stumpf was of the single-stage type, both pressure 
and velocity, like the De Laval, but with this radical difference — a 
wheel about ten times as large in diameter as the De Laval wheel was 
used and, therefore, the same peripheral speed was obtained with 
about one-tenth as many revolutions. The reduction which De Laval 
accomplished by means of gears, Riedler and Stumpf accomplished 




Fig. 43. Milled Buckets in Riedler-Stumpf Turbine 

by increasing the diameter of the wheel. By this reduction in the 
number of revolutions, the error in balance, which, it is claimed, could 
be brought to less than ^J 7 millimeter, was rendered insignificant. 

Their wheels were said to be made of 10 per cent nickel steel 
with 135,000 pounds tensile strength, and were 6| to 9 feet in diam- 
eter, revolving about 3000 r.p.m. for machines of 2000 to 3000 
horsepower. Their single-stage turbine did not meet with general 
favor, and was usually compounded either by pressure stages or 
velocity steps, but a description of it will, nevertheless, be valuable. 

Instead of using vanes of the De Laval type, U-shaped buckets 
were milled in the face of the solid wheel, overlapping one another 
as shown in Fig. 43. The steam jet impinged on the buckets — not 



STEAM TURBINES 



67 



on the side of the wheel, as in the De Laval type, but directly upon 
the face of the wheel — thus permitting a more nearly complete 
reversal of the steam jet and, other things being equal, a higher 
efficiency. It will be recollected that if the jet is delivered to the 
vanes at the side, and at entrance and exit makes an angle with the 



, ] 



n n 





Fig. 44. 20-H. P. Riedler-Stumpf Turbine and Direct-Connected Generator 

plane of rotation, the velocity of the jet* is divided into two com- 
ponents. The velocity of whirl, which is equal to Vcos a, the angle 
a usually being 20 to 35 degrees, is the only component that produces 
a rotative effort. 

The nozzles were made with a square instead of an elliptical 

*Page 17, Part I. 



68 



STEAM TURBINES 



section at the outlet, and were arranged at regular intervals about 
the casing, as in the De Laval turbine. With a given size of wheel, 
the power was increased by increasing the number of these nozzles until 
steam injection took effect upon the entire periphery of the wheel. 

There being only one rotating wheel, it overhung the shaft bear- 
ing, thus passing through the casing on one side only, requiring 
but one stuffing box and, therefore, 
giving a comparatively small bearing 
loss. A 20-h.p. turbine of this type, 
with a direct-connected dynamo, is 
shown in Fig. 44. Fig. 45 shows details 
of the wheel. This wheel is fitted with 
double buckets, which were generally 
used on the large sizes. A 5000-kw. 
turbine of this type would require a 
wheel 20 feet in diameter, admitting 
steam to the whole periphery and mak- 
ing 1500 revolutions per minute. More 
details of the Riedler-Stumpf turbines 
will be described in connection with 
the compound turbine. 



11 




ora 



Fig. 45. Detail of Wheel of Riedler-Stumpf Turbine 

COMPOUND IMPULSE TURBINES WITH VELOCITY STEPS 

It has been shown that steam may be fully expanded to the back 
pressure in a single nozzle, and the kinetic energy absorbed by passing 
the jet through several sets of revolving wheels or vanes in succession, 
each taking out part of the velocity. To employ velocity steps, some 
sort of reversing device must be arranged to bring the steam back, 
either onto another bucket of the same wheel or onto a bucket of an 
adjoining wheel attached to the same shaft. The former method was 
adopted in the Riedler-Stumpf turbine, the latter in the Curtis. In 
either case, a simple and compact turbine is the result, but the 



STEAM TURBINES 



69 



type has disadvantages already enumerated in "Steam Turbines", 
Part I, which limit the number of velocity steps that can be econom- 
ically used to three or four at the outside. Since the work from 
the fifth action of the steam would theoretically be only one-ninth 





Fig. 46. Double Buckets in Riedler-Stumpf Turbine 

of that derived from the first action,* and might easily be consumed 
in additional friction, it is customary to allow the steam to act no 
more than four times. Single-stage turbines are not considered 
practical in sizes above 200 or 300 h.p., it being more economical 



♦Page 31, Part I. 



70 



STEAM TURBINES 



in such cases to reduce the high steam velocity after the first stage 
by using at least one other pressure stage. 

Curtis Turbine. The earlier forms of Curtis turbine were of 
the single-pressure-stage type with several velocity steps, and the 
smaller turbines now made by the General Electric Company are 

after this pattern . Sizes of 35-kw . 
and smaller have a single-pressure 
stage with three velocity steps, 
that is, three sets of rotating vanes 
with two intermediate sets of 
guide vanes. The details of con- 
struction are in all ways similar 
to those of the ordinary form of 
Curtis turbine, which is com- 
pounded both by pressure and 
velocity, and will be described 
under the latter heading. 

Riedler=Stumpf . Large pow- 
ers of the simple impulse type 
required either abnormally large 
wheels or too high speeds of rota- 
tion, and it was, therefore, fre- 
quently more convenient to ex- 
tract the velocity from the steam 
jet in two steps. For powers larger 
than could be dealt with in the 
single-stage type, the steam passed 
successively through buckets of 
the same wheel, and for still larger 
powers, pressure stages were em- 
ployed, as well as the velocity 
steps. The compound- velocity 
turbines developed by Professors 
Riedler and Stumpf had wheels and buckets of the general type 
described in connection with their simple impulse turbine. The device 
employed to reverse the direction of the steam and bring it back 
again to other buckets on the same wheel was described on Page 9, 
Part I, to which the student is referred. 




Fig. 47. Double Guide Vanes 



STEAM TURBINES 



71 



In one type of their turbine the buckets were double, a small 
bucket on one side of the wheel being for initial admission; and, since 
part of the steam velocity was abstracted, it was necessary that, as 
the steam returned, it should enter a larger bucket which was pro- 
vided on the other side of the wheel, as shown in Figs. 46 and 47. 

Another device of Riedler and Stumpf for reducing speeds of 
rotation was the employment of counter-running wheels. The guide 
vanes were buckets cut on a wheel which, instead of being stationary, 
was free to revolve in a contrary direction Thus the absolute 




Fig, 48. Rotor aad ohalt oi Terry Single-Stage Turbino 
Courtesy of Terry Steam Turbine Company, Hartford, Connecticut 



velocity of each wheel would be half the relative velocity of the two 
wheels. In a turbine of this type, besides the obvious objection 
of rotation in two directions, the wheel of initial admission would 
do more work than the counter-running wheel, because the work 
absorbed would be in proportion to the difference of the squares 
of the steam velocities at entrance and exit, and the higher velocities 
would naturally exist in the first wheel.* 

Terry Turbine. The turbine developed and now built by the 
Terry Steam Turbine Company of Hartford, Connecticut, is, in 



♦Page 31, Part I. 



72 



STEAM TURBINES 



sizes up to 1000 h. p., of the single- or two-stage, compound- velocity 
type. The buckets are U-shaped, milled on the face of the wheel, 
overlapping one another something like the single bucket arrange- 
ment of the Riedler-Stumpf machines. Fig. 48 shows the rotor of a 




Fig. 49. Terry Horizontal Turbine with Casing Raised for Inspection of 

Rotor and Bearings 

Courtesy of Terry Steam Turbine Company, Hartford, Connecticut 



Terry turbine. The steam is expanded in the nozzles to the 
exhaust pressure or vacuum. As it leaves the nozzles it impinges 
upon one side of the bucket, reversing through 180 degrees. As 
it leaves the first bucket, it enters a similar bucket attached to 
the casing, which reverses its direction through 180 degrees and 



STEAM TURBINES 



73 



causes it to impinge again upon another bucket of the wheel, and 
so on, until the velocity is all absorbed. The reverse buckets are 
arranged in groups (usually of four), one group for each nozzle, the 
steam being returned to the wheel as many times as there are reverse 
buckets in each group. Fig. 49 clearly shows these buckets on the 
inside of the lifted casing. A crescent-shaped hole may be seen cut 
in the bottom ot each reverse bucket. These holes release a part 
of the expanded steam and thus reduce the volume in proportion 




Fig. 5U. Perry Turbine Jet and Reversing Chambers 

Courtesy of Terry Steam Turbine Company, 

Hartford, Connecticut 

to the lessened velocity, as otherwise there ought to be successively 
larger passage areas. 

In Fig. 50 is shown how a nozzle section, together with its 
reversing buckets, is bolted to the steam chamber of the casing. 
The path of the steam is clearly indicated here. 

There are usually four to eight nozzles in the turbine, each being 
controlled by a hand valve, so that the power may be easily regulated. 
The main bearings are of the ring-oiling type. As the weight of the 
rotor is comparatively small and the speed of revolution 1250 r.p.m. 
for 200 to 300 h.p., large sizes offer no practical difficulty. 



74 



STEAM TURBINES 



Sturtevant Turbine. The B. F. Sturtevant Company of Hyde 
Park, Massachusetts, builds a turbine in small sizes to drive electric 
generators and blowers. In sizes of 100 h.p., or less, these turbines 
have a single-pressure stage, using the steam over and over again on 
the wheel in much the same manner as is done in the Terry turbine. 
Powers of 200 h.p., or over, would be built with two or more pressure 
stages. The wheel is a single forging of open-hearth steel. The 
buckets, which are the U-shape type, are cut from the solid rim 
by a milling machine. 

The earlier turbines have buckets cut on the side of the wheel, 
as shown in Fig. 51. Steam entering the outer edge of these buckets 

passes through the buckets into 
stationary reverse guides in the 
casing shown in Fig. 52. At A 
are inserted the nozzles, which are 
of the ordinary expansion type 
with elliptical openings. The 
guides are of two types ; about four 
are U-shape, like the buckets, and 
return the steam to the wheel 
again, returning it as many times 
as there are return buckets; the 
others, shown at C, are cut open 
at the inner edge in such a man- 
ner that the steam, instead of 
returning to the wheel, is exhausted into the middle of the casing 
and there allowed to pass out. To avoid a troublesome end thrust 
in this machine, buckets are cut on both sides of the wheel, thus 
equalizing the side pressure. 

The Sturtevant turbine as built at the present time is illus- 
trated in Fig. 53. The nozzles and reversing buckets are made of 
Tobin bronze and are cast together in segments. These segments 
are bolted to the casing, which contains an annular steam chamber. 
The nozzle is at A, with four reversing buckets in front of it and 
one supplementary bucket behind it, the purpose of which is to 
utilize any steam which might escape over the back edge of the 
nozzle. These turbines are built in sizes from 5 to 250 horsepower, 
the diameters of the rotors ranging from 12 to 30 inches. 




Fig. 51. Wheel of Sturtevant Turbine with 
Buckets Milled from Side of Wheel 



STEAM TURBINES 



/t> 




Fig. 52. 



Sturtevant Turbine Showing Reverse 
Guides in Casing 




Fig. 53. Parts of Sturtevant Single-Stage Turbine 
Courtesy of B. F. Sturtevant Company, Hyde Park, Massachusetts 



76 



STEAM TURBINES 



De Laval Impulse=Stage Turbine. The class "C" turbines 
manufactured by the De Laval Company are of the velocity-stage 
type. They are built with either two or three velocity stages in 
sizes from 1 to 600 horsepower. Fig. 54 shows a sectional view of 
a three-stage turbine. The steam enters through a set of nozzles 
in which it is expanded to condenser or exhaust pressure. The steam 
then strikes the first impulse wheel. When it leaves this wheel its 
velocity has been reduced by approximately twice the velocity of 




Fig. 54. Axial Section of Class "C" Turbine with Three Velocity Stages 
Courtesy of De Laval Steam Turbine Company, Trenton, New Jersey 

the vanes. It is redirected by a set of stationary buckets to the 
second impulse wheel, and so on to the exhaust end of the turbine. 
The wheels are forged-steel disks keyed to a heavy shaft which, 
unlike that in the single-stage impulse turbine, is entirely rigid. 
The nozzles are of exactly the same type as in the single-stage 
turbine. Both the moving vanes and the stationary guide vanes 
have a crescent-shaped cross section and are of practically the same 
type as is used in the single-stage turbine. The guide vanes are 
held in steel rings which are split on a horizontal diameter, as shown 
in Fig. 55. 



STEAM TURBINES 



77 



Westinghouse Impulse Turbine. The Westinghouse Machine 
Company builds impulse turbines of the re-entry type in sizes up 



Fig. 55. Stationary Guide Vanes in De Laval Turbine, and Method of Mounting 
Courtesy of De Laval Steam, Turbine Company, Trenton, New Jersey 

to 500 horsepower. Fig. 56 shows in diagrammatic form the con- 
struction of this turbine. The steam is expanded in the nozzle; 
it then strikes the crescent-shaped vanes of the rotor, giving up a 




Steam Inlel 



Fig. 56. Section of Westinghouse Re-Entry Type Impulse Turbine 
Courtesy of Westinghouse Machine Company, East Pittsburgh, Pennsylvania 

portion of its energy. It then enters the reversing chamber on 
the opposite side of the wheel and is redirected onto the same set 



78 



STEAM TURBINES 



of moving vanes, giving up more of its energy. The steam now 
enters another nozzle, where it is expanded to exhaust pressure, 
and the operation is repeated. In the small sizes and for pressures 
less than 125 pounds there is only one nozzle and one reversing 
chamber. 

The construction of the nozzle and the reversing chamber is 
clearly shown in Fig. 57, while a sectional view of the turbine is 
shown in Fig. 58. The rotor is a steel disk having a groove in its 




Fig. 57. Construction of Nozzle and Reversing Chamber 
of Westinghouse Re-Entry Type Impulse Turbine 
Courtesy of Westinghouse Machine Company, 
East Pittsburgh, Pennsylvania 

periphery. The vanes have a shank at the root fitting into the 
groove and are riveted to the rotor. The outer end is fitted with 
shroud ring. 

COMPOUND IMPULSE TURBINES, PRESSURE STAGES 

In the discussion of compound turbines in Part I it was shown 

that the available head could be divided into several stages, thus 

making the steam velocity from stage to stage relatively small, and 

permitting smaller speeds of revolution. Turbines of this type are, 



STEAM TURBINES 



79 



in principle, like a number of De Laval wheels on the same shaft. 
They consist essentially of a casing which supports a number of 
diaphragms, dividing the interior into separate cells, in each of which 
a single impulse wheel containing the vanes is free to revolve. Each 
stage or element comprises a rotary wheel and a set of nozzles, or 
distributing vanes, which guide the steam from one chamber to the 
next and direct it at the proper angle onto the vanes of the wheel in 
the following chamber. These passages may or may not be of the 
diverging type, depending upon the drop of pressure from stage to 




Fig. 58. Sectional View of Westinghouse Re-Entry Type Impulse Turbine 
Courtesy of Westinghouse Machine Company, East Pittsburgh, Pennsylvania 

stage. In all machines of this type the drop in pressure is so arranged 
that an equal number of heat units will be given up per stage, which, 
as will be remembered, does not correspond by any means to an equal 
drop in pressure. 

In such a turbine as this a foreign substance is not likely to 
injure more than one wheel, for it cannot pass the diaphragm separa- 
ting the different chambers except through the nozzles and, as 
there are many stages to such a turbine as this, the machine might 
run fairly well, even if one or two wheels were removed. It would, 



80 



STEAM TURBINES 



of course, give less power and poorer steam economy. The clearance 
between the nozzles and the vanes should be small to prevent the 
mingling of the steam jet with the stagnant steam in the casing, but 
the clearance over the ends of the vanes is of little consequence, 
especially if a shrouding is used, for there is no tendency for steam to 
leak by the vanes, the pressure being constant throughout the chamber. 




Fig. 59. Group of Diaphragms of Rateau Turbine 

Rateau Turbine. The turbine using pressure stages, only, is 
best exemplified by the Rateau turbine, designed and developed by 
Professor Rateau, of Paris. His turbine is a horizontal, multi-stage 
impulse machine and consists of sometimes as many as forty pressure 
stages, but usually less. The large number of stages employed in 
these turbines means but little drop in pressure from stage to stage. 
Hence, the De Laval type of expanding nozzle is not needed, it 
having been shown in the discussion of nozzles* that, if the final 

*Page 19, Part I. 



STEAM TURBINES 81 

pressure exceeds 58 per cent of the initial pressure, a parallel-sided 
passage or a slightly converging nozzle is sufficient to permit the 
proper expansion of the steam and to secure the maximum available 
energy. In the Rateau turbine, the drop in pressure from stage to 
stage is much less than the limiting amount mentioned. 

Diaphragm, As the steam passes through the turbine, it 
expands from one stage to the next, and, of course, requires larger 
passage areas in each succeeding diaphragm. This is accomplished 
in general by . increasing the number of openings rather than by 
increasing the size of them. The guide passages are arranged in 
groups, the number in each group being increased and, consequently, 




Fig. 60. Lower Half of Turbine Casing Showing Dia- 
phragms in Place, and the Circumferential Grooves 
for Holding Them 

the width of the group widened through successive stages until the 
openings finally extend entirely around the disk. To provide larger 
passage areas to take care of still further expansion of the steam, the 
diameter of the wheel and diaphragm must be increased and, at the 
same time, the size of the passage openings enlarged. In condensing 
turbines there are usually three diameters of wheel. In non-con- 
densing turbines, two are generally sufficient, and sometimes only 
one is used. The nozzles, or distributors, when only a portion of 
the wheel is open to steam admission, are set to have an angular 
advance of the preceding group, this advance being proportioned to 
the speed of the wheels, so that the steam jet as it leaves the revolv- 
ing vanes will strike the next nozzle directly, avoiding any shock of 



82 STEAM TURBINES 

impact against the solid wall of the casing. Any kinetic energy 
in the steam as it leaves the revolving vanes of one wheel is, therefore, 
directly available for use in the next stage. If the steam were 
brought up sharply against the solid casing wall, this residual energy 
would be lost to useful work, and a still further loss would result, 
due to the eddying of the steam in this particular cell. Fig. 59 
shows a group of diaphragms of various sizes for this turbine and 
illustrates the idea of increasing the extent of each succeeding group 
of nozzles. 

There is a distinct advantage in this partial admission at the 
higher pressures, for, if the admission took place around the entire 
periphery of the wheel, the height of vane would necessarily be so 
small that the friction would be excessive. By using partial admis- 
sion only, the vanes in these stages may be of much greater height 

than otherwise, and a few 

^^^HB^^^^^^S^^^S. n ^ n vanes afford the same 

^Sgs • ^SSHSjjk * Jm passage area for the steam 

^S& f ^VT\ !« that a large number of low 

fS^? '£k ' A^^ ^ *'m vanes would offer, with con- 

* ff J sequently less friction. 

'* +immn | n ^k ImP'; Casing. Circumferen- 

•* > " > mKm tial grooves are turned in 

Fig. 61. Two-Piece Stationary Diaphragms with the inside of the Casing to 

Distributing Vanes . 

hold the diaphragms in 
place, as shown in Fig. 60. The larger diaphragms are usually made in 
two pieces, as shown in Fig. 61 . In all of them the shaft passes through 
collars of antifriction metal with clearances as small as possible, in 
order to prevent leakage from stage to stage. These collars are fre- 
quently provided with a labyrinth packing which will be explained later. 

Vanes. The vanes are of crescent shape, similar in cross section 
to those of the De Laval turbine. The earlier ones had no protection 
at the outer ends, but the later ones have been provided with a shroud 
ring to give additional stiffness. The vanes and the shroud ring 
are generally made of nickel steel, as this is well adapted to resist 
corrosion. 

Wheels. The rotating wheels are usually made of two plate- 
steel disks. One is flanged at the outer edge to which the vanes are 
riveted, the other, usually slightly conical, is riveted to the flange of 



STEAM TURBINES 



83 



the first disk, and both are riveted to the cast-steel hub. Fig. 62 
shows the construction of this type of disk. The conical built-up disk 
makes a stronger wheel, but a flat disk is sometimes used. Some 
wheels are turned out of solid steel, increasing in section toward the 
center to give greater strength. Each wheel is carefully balanced by 
itself on knife-edges by drilling holes in the disk, the latter clearly 
showing in Fig. 63. When these wheels are assembled on the shaft 
there is little likelihood that the complete rotor will be out of balance. 
Fig. 64 shows the wheels assembled on the shaft. 




Fig. 62. Construction of Wheels of the Rateau Turbine 

By-Pass Valve. This turbine is provided with a by-pass valve 
to carry overloads. This admits high-pressure steam into the inter- 
mediate stages and, although not permitting a complete expansion of 
such steam, it is an effective means of taking care of large overloads. 

Bearings. The bearings are of the plain ring-oiling type, usually 
provided with water jackets. The shaft not being unduly long, there 
is little danger of whipping, and as the speeds of rotation are not very 
high, special precautions are not necessary. Sometimes the turbines 
are supported by three bearings, as shown in Fig. 65, the high and 
intermediate stages being separated from the low by a third bearing. 



84 



STEAM TURBINES 



Professor Rateau claims for a 1500-horsepower turbine at 1500 
revolutions per minute only 1| per cent leakage and bearing loss, 
and a 2J per cent loss in friction of the wheels against the steam. 

Zoelly Turbine. The Zoelly turbine has been developed 
extensively abroad, and is being manufactured largely through a 
syndicate of builders including some American firms. It is a turbine 




Fig. 63. 



Wheels of Rateau Turbine, Balanced by 
Holes in Disk 



essentially of the Rateau type, that is, a multi-pressure-stage turbine, 
but it has fewer stages, usually not over ten for condensing and five 
for non-condensing, and it differs from the Rateau materially in detail. 
The blades are very much longer, sometimes being as much as half the 
diameter of the rotating wheel and, because of the fewer stages, many 
essential details are different. The turbine is sometimes divided into 
two parts when built condensing, the high pressure being separated 



STEAM TURBINES 85 

from the low sufficiently for a third bearing to be placed between, as 
shown in Fig. 66. 

Vanes. There being fewer stages than in the Rateau, the steam 
velocities must be much greater and, consequently, if the rotative 
speeds are to be the same, the diameter of the wheels must be greater. 
Exceedingly long vanes are used, which permit of a relatively small 
wheel disk; hence, the centrifugal stresses in its rim will not be 
materially greater than in other turbines. Again, the vanes are few 
in number compared with other turbines and are made as light as 
possible, tapering at the outer end in order to further reduce the 
centrifugal force. The expansion from stage to stage is not enough 




Fig. 64. Shaft and Wheels of High-Pressure Rateau Turbine 

to require diverging expansion nozzles except, perhaps, in the 
high-pressure end; but the expansion is, of course, very much greater 
than in the Rateau turbine. In the latter turbine, the roots of the 
vanes are cut off parallel with the shaft, but in the Zoelly they are 
cut on a slope, giving a larger outlet than inlet to the vane. It might 
appear that this is done to permit of steam expansion in the vanes, 
but that is not so. Expansion is complete in the nozzles just as in 
the Rateau turbine, but, as there is a much greater expansion from 
stage to stage, it follows that the area of the steam passages through 
each successive diaphragm must be greatly increased. This differ- 
ence is made up by increasing the depth of the openings in the 
succeeding diaphragm. Therefore, to permit a free passage of steam 
from an opening having one depth to an opening having a greater 



88 



STEAM TURBINES 



depth, and to prevent the formation of eddies, it is necessary to slope 
the root of the vanes. The vanes of the turbines being very much 
longer than in the Rateau, it is entirely feasible to increase the depth 
of the steam passages through the diaphragms. 

Wheel. The vanes are set in slots cut in the rim of the wheel, 
and are secured by a clamp ring securely riveted to the main portion 
of the wheel, as shown in Fig. 67. It is probable that the excessively 
long vanes produce a considerably greater friction loss, revolving, 

as they do, in the steam-filled cells 
of the casing; but as there are com- 
paratively few cells, and therefore 
comparatively few revolving wheels, 
it is probable that this friction may 
not be any greater than would be 
expected in the Rateau turbine. 
Economies, as shown by test, do not 
appear to be essentially different 
from those of other first-class tur- 
bines. Zoelly turbines have been 
built as shown in Fig. 68. The 
typical style of long vanes prevails, 
but in the high-pressure stages the 
vane is of the double-U-shaped cross 
section, a detail of which is shown 
in Fig. 69. The steam jet necessarily 
impinges tangentially on these vanes 
instead of from the side. 

Hamilton=Holzwarth Turbine. 
The Hamilton-Holzwarth steam tur- 
bine is another turbine of the Rateau 
type, the chief difference being that, instead of having the admission 
guides arranged in groups, the admission of steam in the high- 
pressure end takes place around the entire circumference of the 
diaphragm. Hence, the wheels of this turbine would be smaller at the 
high-pressure end than would be the case with the Rateau turbine, and 
the vanes would be appreciably of less depth. Theoretically, the diam- 
eter of each succeeding wheel should increase as the steam expands but, 
for simplicity of manufacture, it is better to keep a number of wheels 




■Section C-C 

Fig. 67. Detail of Wheel and Vanes 
of Zoelly Turbine 



90 



STEAM TURBINES 



of the same diameter, increasing the length of blades to give larger 
passage areas. When the point is reached where this is no longer 
practicable, the diameter of the wheel may be increased considerably 
and the depth of the blade reduced. 

There are approximately the same number of stages in the 
Hamilton-Holzwarth turbine as in the Rateau, and the speeds of 




69. Details of Construction of the Double-U-Shaped 
Vanes of the Zoelly Turbine 

revolution are not materially different. The running wheel shown in 
Fig. 70 consists of a steel disk riveted to each side of the steel hub, or 
spider, which is keyed to the shaft. The outer edges of the disk are 
flanged outward, leaving a space to take the shank of the vane. The 
vanes are drop-forged steel of the usual crescent-shaped cross section. 
The shank of the vane, fitted between two flanged disks, is riveted 



STEAM TURBINES 



91 



in place. Since the steam is admitted around the whole circum- 
ference of the diaphragm, vanes can be used better than nozzles to 
give the necessary expansion and direct the steam upon the running 
wheels. Vanes could be used in the Rateau diaphragm even with 
partial admission, but in such a case it has seemed simpler to drill 
openings through the diaphragm. 





=3 



Fig. 70. Details of Wheel of Hamilton-Holzwarth Turbine 



For small-pressure drops the vanes may be parallel, top and 
bottom, but, at the low-pressure end of the turbine where the volumes 
increase rapidly, they are usually deeper at the outlet than at the 
inlet, thus forming an easy passage for the steam to the next larger 
set of vanes. 

A diaphragm, Fig. 71, separating the various cells, consists of a 
cast-iron disk bored loosely to fit the shaft. This disk has a groove C 
into which the shanks of the guide vanes are set, and a rivet holds 
them in place. A steel band is then shrunk over the ends of these 



92 



STEAM TURBINES 



guide vanes, and this band projects into grooves in the casing to hold 
the diaphragm in place. 

The Hamilton-Holzwarth turbine was manufactured by the 
Hooven, Owens, Rentschler Company of Hamilton, Ohio. It is 
not being built at the present time. 

De Laval Pressure=Stage Turbine. To meet the demand for 
large-power and slow-speed turbines the De Laval Turbine Company 




Fig. 71. Details of Diaphragm of Hamilton-Holzwarth Turbine 

have recently developed a turbine of the pressure-stage type. The 
general arrangement is shown in Fig. 72. The rotating member 
consists of a heavy shaft upon which is mounted a series of disks, 
each revolving in its own chamber, formed by diaphragms held in 
the cylindrical casing. The steam is admitted to the first stage 
through a number of cylindrical nozzles formed in a nozzle ring 



STEAM TURBINES 



93 



which separates the steam chest at the right-hand end of the casing 
from the first wheel. 

The vanes are of the same construction as those used in their 
single-stage impulse turbines, increasing in size progressively toward 
the exhaust end of the turbine. 

With the exception of the nozzles in the first stage, the dia- 
phragms separating the succeeding stages have guide vanes formed 
around their entire circumference. The vanes are located on the rim 
of the diaphragm by means of two pins for each vane, and then a 




Fig. 72. 



Axial Section Showing General Arrangement of De Laval Multi-Stage Steam Turbine 
Courtesy of De Laval Steam Turbine Company, Trenton, Xew Jersey 



steel band is shrunk over them to hold them in place. These bands 
project beyond the sides of the vanes and form a strong lining for 
the cast-iron casing, as shown in Fig. 73. 

Wilkinson Turbine. The James Wilkinson Company developed 
a turbine of the multi-cellular impulse type with comparatively few 
stages, which is noteworthy because of the packing employed to pre- 
vent leakage of steam from stage to stage. It must be borne in 
mind that in a turbine of the Rateau type there are so many stages 
that there is comparatively little difference in pressure between one 



94 



STEAM TURBINES 



stage and the next, and consequently little tendency for steam to 
leak through between the shaft and the bushing in the diaphragm. 

As the number of stages is 



reduced, the difference in 
pressure is increased, and it 
then becomes essential that 
some sort of packing should 
be provided; otherwise the 
leakage may be excessive. 
In practically all turbines, 
to prevent leakage of steam 
from the high-pressure end 
into the air, a labyrinth 
packing is used, consisting 
of a series of grooves into 
which metal spring rings are 
fitted with slight clearances. 
These rings are not very 
different from the spring 
rings employed to pack the piston of the reciprocating engine, except 
that there are usually very many more of them. Steam, in order to 
leak out, must follow a tortuous course between the rings and sides 




Fig. 73. Partial Section of De Laval Multi-Stage 
Turbine Showing Steel Retaining Rings 




Fig. 74. Grooved Steam Passages and Labyrinth Packing of Wilkinson Turbine 

and bottom of the grooves, greatly expanding and condensing as it 
leaks through. The volume at the outer end is so large, due to this 
expansion, that the leakage becomes very slow. Condensation from 
such a packing is usually caught by a drip and taken to the hot well. 



STEAM TURBINES 



95 



The Wilkinson idea is to groove passages along the shaft between 
certain rings of the labyrinth packing and the diaphragm bushings, 
as shown in Fig. 74. These grooves permit steam to pass to the 
diaphragm bushing from a groove in the labyrinth packing, which 
is at slightly higher pressure than that in the cell at either side of the 
diaphragm bushing. As a result steam will leak from the labyrinth 
packing to the diaphragm bushing and, as the pressure of the steam 
in the cells is less than of that which enters the bushing, this wet 
steam will expand, thereby vaporizing a portion of its moisture, and 
will then leak through into the cells of the turbine and have a possible 
chance of doing some useful work in the lower stages. The leakage 
from the labyrinth packing will be somewhat augumented by this 
means, but the leakage from stage to stage will be practically 
eliminated. 

The Wilkinson turbine has a unique governing device but is 
essentially of the Rateau type of turbine, differing only in detail. 

Kerr Turbine. The turbine as originally built by the Kerr 
Steam Turbine Company of Wellsville, New York, is of the multi- 
stage type and differs from the 
Rateau in three important par- 
ticulars. There are compara- 
tively few stages; the vanes and 
buckets of the rotating wheels are 
of the double cup-shape sort, 
after the style of buckets of the 
Pel ton water wheel; and the steam 
is directed onto the wheel through 
nozzles instead of guide passages, 
striking the wheel tangentially in 
the plane of rotation instead of at 
the side. Otherwise the differ- 
ence between this turbine and others of the multi-stage type is 
one of detail of construction. 

The turbines are built in standard sizes, with wheels 12, 18, 24, 
and 36 inches in diameter, with one to eight stages, developing up to 
600 horsepower. The power developed in a single wheel of given 
diameter will naturally depend upon the number of nozzles in action. 

The buckets are made of drop forgings and are reamed out with 




Buckets of Kerr Turbine 



96 



STEAM TURBINES 



a special reamer, each cup being, in cross section, a surface of revolu- 
tion. Fig. 75 illustrates this type of bucket. 




Fig. 76. Kerr Turbine Wheels Assembled on Shaft 

The wheel is a solid disk and the buckets may be attached by 
riveting, as shown in Fig. 75, or by dovetailing, as shown in the same 




Fig. 77. Nozzles, Wheel, and Diaphragm of Kerr Turbine 

figure, the latter being the preferred and more usual form of construc- 
tion. Each wheel is carefully balanced and then assembled on the 
shaft, as shown in Fig. 76. 



STEAM TURBINES 



97 




98 



STEAM TURBINES 




Fig. 79. Diaphragm of Economy Turbine, Showing 

Nozzle Construction. Left Half, Outlet Side; 

Right Half, Inlet Side 

Courtesy of Kerr Turbine Company, 

Wellsville, New York 




Fig. 80. Section of Bucket Wheel of Economy Turbine, 
Showing Method of Inserting and Holding Blades 



The nozzles are of 
machine steel, screwed 
into a steel nozzle body, 
which is securely riveted 
to the diaphragm casing. 
Fig. 77 : * shows the noz- 
zles, wheel, and dia- 
phragm. 

The shell is cast in 
sections, one for each 
stage. These are set be- 
tween two end casings, 
turned and bored. 
Tongue-and -groove 
joints on these castings 
insure correct alignment, 
and fibrous packing in the 
grooves, in addition to 
metal contact of the sur- 
face, insures steam tight- 
ness. A casing built in 
this manner possesses 
two distinct advantages 
over a solid casing. There 
is no probability of crack- 
ing due to rapid temper- 
ature changes, and the 
size of turbine may be 
increased by the addition 
of more units. The end 
castings carry the weight 
of the turbine and have 
supports bolted to the 
bed plate. 

The shaft, where it 
passes through the dia- 
phragm, is fitted to a 
bronze bushing with a 



STEAM TURBINES 99 

few thousandths of an inch clearance. This bushing seats on the 
metal surface of the diaphragm with latitude for slight side motion. 
It is kept to its seat by the steam pressure, but can move sideways 
to accommodate any whipping motion of the shaft. The turbine 
as built at the present time by the Kerr Company is quite different 
from the original type. A sectional view of it is shown in Fig. 78. 
The Pelton form of bucket has been abandoned as well as the separate 
nozzles. These are formed by walls within the diaphragms and thin 
metal vanes die-pressed into shape and cast into the diaphragm. They 
are of slightly converging form. Fig. 79 shows a diaphragm with 
nozzles. 

The buckets as now made are shown in Fig. 80. They are drop 
forgings held in dovetail slots and riveted. The outer ends are 
rigidly held in a shroud ring to which they are riveted. 

COMPOUND IMPULSE TURBINE WITH PRESSURE STAGES 
AND VELOCITY STEPS 

One of the simplest and most effective ways of compounding 
turbines is by both pressure stages and velocity steps. The turbine 
shell is divided by diaphragms into a number of different cells, seldom 
more then five, except for marine work, where more are necessarily 
employed. There being comparatively few stages, it will, in most 
cases, be necessary to employ diverging nozzles so proportioned that 
the steam may be completely expanded within the confines of these 
nozzles, from the initial pressure to the pressure in the chamber into 
which the steam is discharging. It will be remembered that multi- 
stage turbines of the Rateau type do not require expanding nozzles 
because of the relatively small drop in pressure from stage to stage. 

Turbines of the type now being described differ from the Rateau 
type in another most important particular. Each cell, or chamber, 
of this type contains two or more sets of rotating vanes, while turbines 
of the Rateau type have but one wheel and one set of vanes in each 
chamber. The steam on leaving the nozzles impinges upon the first 
set of running vanes, and, as the steam leaves these vanes, it flows 
into a set of guides of some sort and, as the case may be, is returned 
to other vanes of the same set, or to a different set of vanes on the 
same wheel, or to the vanes of a separate wheel. The steam may 
pass through from one to three sets of redirecting guides, and may 



100 



STEAM TURBINES 



fc £ •* £ 
Jill 



it 
I? 




Fig. 81. Plan and Elevation Showing Steam Passages in Curtis Two-Stage Steam Turbine 
Courtesy of General Electric Company, Schenectady, New York 



STEAM TURBINES 101 

impinge upon two to four sets of rotating elements. It is immaterial, 
so far as the principle of the action is concerned, whether the steam 
acts successively upon a number of rotating wheels or whether it is 
returned again and again into different vanes of the same wheel. 
If the latter form of construction is adopted, the turbine will neces- 
sarily be more compact and the rotating shaft will be shorter. 

The more velocity steps the turbine has per stage, the fewer 
number of stages will be necessary, but, in general, it is found more 
economical to increase the pressure stages than to increase the 
velocity steps. It must be remembered that in this type of turbine 
the steam is completely expanded within the nozzle, and that the 
temperature and pressure of this expanded steam are the same within 
the confines of any particular cell. As the steam passes through 
successive rotating vanes, it gradually loses velocity and, conse- 
quently, the succeeding passages must be made larger and larger, 
in order that the same volume of steam may pass through at the 
lower velocity in a given time. In other words, the passage area 
must increase in proportion to the reduction in velocity. If this 
point is not clearly borne in mind when looking at the vanes of a 
turbine of this type, one might think that the increased size of the 
passages was to provide for increased expansion of the steam. In 
the Terry steam turbine, which is of this type, a portion of the steam 
is allowed to escape as it passes through successive buckets, so that 
the volume is gradually reduced as the velocity decreases. 

Curtis Turbine. Undoubtedly the best-known turbine of this 
type is the Curtis, which has been developed by the General Electric 
Company, and is being built at their works at Schenectady, New 
York, and Lynn, Massachusetts. Rights to build the Curtis turbine 
for marine propulsion are controlled by the Fore River Ship-Building 
Company of Quincy, Massachusetts. 

For ordinary purposes, up to 9000 kilowatts, the Curtis turbine 
does not have over four stages, with two velocity wheels per stage, 
except in the larger sizes of turbine, when a fifth cell is provided, 
which contains a single rotating wheel without redirecting guides, 
for abstracting any residual velocity there may be in the steam after 
passing the fourth stage. The marine turbine of this type installed 
in the United States Cruiser Salem had seven pressure stages with 
four velocity steps in the first stage and three in each of the others. 



102 



STEAM TURBINES 



Fig. 81 shows in diagrammatic form the principle of the steam 
action in the Curtis turbine. Steam passes from the steam chest 




through the nozzles, each set of which may be closed by a valve 
operated from the governor. The number open at any one time 
depends upon the load. There are enough valves and nozzles to 



STEAM TURBINES 103 

take a large overload without the use of the by-pass, which would 
admit high-pressure steam into the lower stages. Such a by-pass is, 
however, usually provided. It works automatically and admits 
high-pressure steam to a set of auxiliary nozzles fitted into the second 
stage of the turbine, thus increasing the power with only a slight 
sacrifice in steam expansion and consequent economy. 

The nozzles are designed to produce such a drop in pressure from 
stage to stage that equal amounts of work are done in each stage. 
This does not correspond to an equal drop in pressure by any means, 
for there are more heat units in a given drop of pressure in the lower 
than in the higher ranges. With many stages in a Curtis turbine, the 
low-pressure diaphragms might be fitted with plain cylindrical nozzles. 

Steam enters the first row of rotating vanes* from the nozzles, 
is deflected from these vanes to the guides, or intermediates as they are 
technically called by the General Electric Company, and is redirected 
to the next row of moving vanes, and so on, passing from the last 
row directly into the next and again to the next stage as before. 

As has already been mentioned, the number of pressure stages 
and velocity steps in the Curtis turbine vary with the size of the unit. 
In sizes from 75 to 300 kilowatts, there are two stages and three 
velocity steps, as shown in the diagram, Fig. 81. The 500-t to 
3000-kilowatt sizes are four-stage, with two velocity steps per stage, 
while those of over 3000 kilowatts are five-stage, with only a single 
wheel in the fifth chamber and, of course, no reducing buckets. 
The turbines, in sizes up to 300 kilowatts, are generally of the hori- 
zontal type, the larger sizes being vertical. Fig. 82 shows a 100- 
kilowatt horizontal turbine, and Fig. 83, a vertical turbine of 9000- 
kilowatt capacity. A sectional view of the turbine as fitted in the 
U. S. S. Salem is shown in Fig. 84. 

It may be interesting to note the speeds of rotation of Curtis 
turbines of various sizes. 



500 kilowatts 


approx. 


1800 r.p.m. 


1000 kilowatts 


approx. 


1200 r.p.m. 


2000 kilowatts 


approx. 


900 r.p.m. 


5000 kilowatts 


approx. 


800 r.p.m. 


9000 kilowatts 


approx. 


750 r.p.m. 


Marine turbines 


approx. 


250 r.p.m. 



*The rotating vanes are called buckets by the General Electric Company. 
fThe General Electric Company builds a special 500-kilowatt vertical turbine having 
only two stages with, consequently, three rows of revolving vanes per stage. 




to 



1\ Q 71 °^ 



^ig. 83. Vertical Curtis Turbine of 9000-Kw. Capacity 



106 



STEAM TURBINES 




Fig. 85. Section of Casing of Four-Stage Curtis Vertical Turbine 



STEAM TURBINES 



107 




Fig. 86. Group of Nozzles for High-Pressure End 



Casing. Fig. 85 shows the section of the casing of a four-stage 
turbine. It is built of cast iron of four parts for sizes up to 3000-kilo- 
watts, and six parts for 5000 kilowatts and larger sizes. This casing 
holds the stationary reversing vanes and supports the diaphragm, the 
details of which are clearly 
shown in Fig. 85. A rep- 
resents the inlet nozzle for 
the first stage, B the guides 
or intermediates, C the 
diaphragm separating the 
first and second stage, D 
the ledge on the casing which supports the diaphragm, E the spider 
of the rotating wheel, F the wheel plates, and G the distance piece 
at the outer rim of the wheels. 

Nozzles. The nozzles are grouped together, not as in the Rateau 

turbine, but in one single group for 
each stage, thus admitting a single 
steam belt to a part of the wheel pe- 
riphery only. The Rateau turbine also 
admits steam to only a portion of the 
periphery, but in this turbine there are 
several groups of nozzles instead of 
one, arranged at equal intervals around 
the periphery. 






Fig. 87. Portion of Diaphragm, 
Showing Construction of Buckets 



Fig. 88. Buckets of 1000-Kw. Curtis 
Steam Turbine 



Fig. 86 shows a group of nozzles for the high-pressure end of 
the turbine. The outlet is of rectangular section, slightly rounded 
at the corners, which the makers claim gives better results with this 



108 



STEAM TURBINES 



type of turbine than the elliptical outlet used in the De LavaL The 
nozzles are reamed out of bronze castings and riveted to the casing. 

Vanes. The vanes, usually called buckets by the Curtis 
manufacturers, are crescent-shaped in cross section, as is common for 
impulse turbines. 

The construction of buckets and wheels is shown by Figs. 87 
and 88. The wheel is a steel disk with the rim enlarged so that a 
groove of dovetail section may be cut in its periphery. The buckets 
have a corresponding dovetailed root which fits snugly into the groove 
of the wheel. At intervals the groove has openings for the insertion 
of buckets, Fig. 87. These openings are then closed by means 




Fig. 89. Rotating Wheels of the Curtis Steam Turbine, 
Showing Two Sets of Vanes 

of a spacing block. After the buckets are assembled, the shroud 
ring is riveted to their outer ends by means of the small projections 
shown on the buckets. This ring serves the twofold function of 
reducing the vibration of the buckets and preventing the jet of steam 
from spilling over their ends. 

Wheels. Fig. 89 shows the rotating wheels of a Curtis turbine, 
with two sets of vanes so spaced that, as the wheel rotates, there will 
be one set of vanes on each side of the guide vanes, or intermediates. 
This is equivalent to having two rotating wheels, but in construction 
is much simpler. Each wheel, as shown, rotates in a chamber by 
itself. Figs. 89 and 90 show the old style of bucket segments which 
offer a very rough surface. In the present style, Fig. 87, the wheels 



STEAM TURBINES 109 

have no external webs and no riveting. The wheels, except in the 
larger sizes, are made of solid steel, securely keyed to the shaft at the 
hub. This hub varies in length according to the stage in which it 
revolves, the low-pressure stages necessarily being wider than the 
high-pressure stages. The wheel gradually tapers toward its periph- 
ery, thus maintaining a section of approximately uniform strength. 
Wheels for the largest sizes of turbine are built up of a cast-steel 
hub, or spider, to which are riveted steel disks, one on each side, and 
between them is riveted a cast-steel distance ring. The vanes are 
riveted in segments onto the outer edges of the disks, as is clearly 
shown in Figs. 83 and 84. 

Guides. The guide vanes, or intermediates, are necessarily in 
one group of sufficient extent to catch the steam belt as it issues from 
the rotating wheels. The group naturally extends at first over only 
a small arc of the circumfer- 
ence, but this arc increases 
in extent in each stage until 
both nozzles and interme- 
diates entirely surround wheel 
and casing. The guide vanes 
are in cross section like the 
revolving vanes, but are set „ f * eam T u i)T. n 

° ' m t Courtesy of General Electric Company 

in a reverse position. They 

are set with a certain angular advance from the nozzles, depending 
upon the speed of rotation, so that the steam leaving the vanes will 
strike full upon them. Like the vanes, they have to provide a 
successively increasing passage area to allow for the lower velocity 
of the steam, if more than one set of guides is used per stage. 

Diaphragm. The style of diaphragm is best shown in Figs. 84 
and 85, which also clearly show the construction of the large wheels. 
The diaphragm is an iron casting — steel in the higher stages because 
the drop in pressure is greatest there, and, consequently, a greater 
load on the diaphragm — slightly dished in shape toward the high- 
pressure side to give greater strength. It is provided with bronze 
bushings where the shaft passes through, these bushings being fitted 
with slight clearance to prevent leakage. The ends of the casing are 
packed with carbon packing to prevent leakage of steam into the 
air at the high-pressure end, and leakage of air into the turbine at the 




110 STEAM TURBINES 

low-pressure end, where leakage would have a detrimental effect upon 
the vacuum. 

Bearings. All sizes of the Curtis turbine of 500 kilowatts and 
over are of the vertical type with only one working bearing, located 
at the lower end of the shaft*. This bearing consists of a short cylin- 
drical block of cast iron fitted with two dowel pins and a key, as 
shown in Fig. 91, and a corresponding cast-iron block having a hole 
through its center, into which a pipe is threaded for supplying some 
form of lubricant, either water or oil. 

Fig. 92 shows the bearing assembled on the lower end of the 
shaft. The upper bearing with dowel pins and key fits into corre- 
sponding dowel holes and keyway in the bottom of the shaft, and 




Fig. 91. Bearing Surfaces in Step Bearing of Curtis Turbine 
Courtesy of General Electric Company, Schenectady, New York 

rotates with it. When the oil is supplied to the bearing, which is, 
of course, under a high pressure, it fills the central circular space 
between the blocks and forces them slightly apart. The oil then 
escapes between the annular edges of these two blocks and is col- 
lected into a drain and returned to the original supply. If water is 
used for a lubricant, it is allowed to flow up into the base of the tur- 
bine and mingle with the exhaust steam on its way to the condenser. 
The pressures maintained by the lubricating pump in practice vary 
from 180 to 450 pounds per square inch. It is thus seen that the 
two bearing blocks do not come into actual contact, but that the 
weight of the turbine is supported upon a film of lubricant. Should 
the lubricating pump fail in its supply, no more serious damage would 



*The horizontal construction is favored at the present time and the General Electric Com- 
pany build all sizes of turbines in that style. 



STEAM TURBINES 



111 



occur than the abrasion of the step bearing, and a new one could 
readily be inserted, as the figure will show. 

Riedler=Stumpf Turbine. The turbine developed by Professors 
Riedler and Stumpf for the larger powers necessitating lower speeds 
was provided with two to four pressure stages, with two velocity 
steps per stage, and this turbine expanded the steam on the same 
principle as the Curtis, but the details of construction and general 



1~-5feam 




Pra/n. 



Oft Suppty, 

Fig. 92. Step Bearing of Curtis Turbine 

arrangement were entirely different. In the two-stage, and even in 
the four-stage, turbine the overhung type of wheel developed in the 
single-stage turbine was adhered to, the generator being between 
the two turbine wheels with a bearing at each end. Fig. 93 shows 
this arrangement in a 5000-kilowatt, two-stage turbine revolving at 
750 r.p.m., with two velocity steps to each stage. In this particular 
turbine, the steam is returned through U-shaped guide passages to a 



112 



STEAM TURBINES 



second set of buckets on the same wheel, these buckets being larger 
than the first because of the lower velocity. Fig. 94 shows a 500- 
kilowatt, four-stage turbine of the same type. These have been 
built both vertical and horizontal, the vertical arrangement resem- 
bling externally the Curtis turbine. 

Fig. 95 shows a vertical four-stage, two-step turbine of 750 
r.p.m. developing the same power as that shown in Fig. 94. 



f& 




Fig. 93. Overhung Wheels of the Riedler-Stumpf Turbine 

The Reidler-Stumpf turbine was formerly manufactured by the 
Allgemeine Elektricitats Gesellschaft of Berlin, but, as this company 
is now licensed to manufacture under Curtis patents, it does not 
appear that the manufacture of the former type is being actively 
carried on at the present time. 

Terry Turbine, In the medium sizes of turbines, up to about 
360 kilowatts, which are to operate condensing, the Terry steam 







Fig. 95. Vertical, Four-Stage, 500-Kw. Riedler-Stumpf Turbine 



STEAM TURBINES 



115 




Fig. 96. 



Section Through Two-Stage, Condensing 
Terry Steam Turbine 




Fig. 97. Section of Return-Flow Terry Steam Turbine 
Courtesy of Terry Steam Turbine Company, Hartford, Connecticut 



116 



STEAM TURBINES 



turbines are of the two-stage type, as illustrated in Fig. 96. The 
turbine consists essentially of two single-stage wheels in the same 
casing, separated by a diaphragm. The nozzles of the first stage 
expand the steam to about atmospheric pressure and, after passing 
through the wheel, it is further expended to condenser pressure in 
the nozzles of the second stage. 




Fig. 98. Return-Flow Steam Turbine with Cover Lifted 
Courtesy of Terry Steam Turbine Company, Hartford, Connecticut 



More recently the Terry Steam Turbine Company have devel- 
oped the turbine shown in Fig. 97. This, as will be seen, consists 
of a Terry turbine wheel in combination with a Rateau pressure 
stage element, the Terry wheel being at the high-pressure end of the 
machine. An interesting feature of this turbine is the so-called 
return-flow principle adopted in its construction. The steam, after 



STEAM TURBINES 



117 



having passed through the Terry wheel, is at slightly above atmos- 
pheric pressure. It now travels to the other end of the turbine and 
passes through the Rateau element in the reverse direction. The 




Exhaust 



Fig. 99. Phantom View of Spiro Turbine, Showing Extreme Simplicity and Compactness 
Courtesy of Buffalo Foroe Company, Buffalo, New York 

purpose of this is to prevent leakage of air into the condenser where 
the shaft passes through the casing. Fig. 98 shows an external view 
of this turbine with cover lifted. 

Buffalo Forge Spiro Tur= 
bine. The Buffalo Forge Com- 
pany have developed a novel 
construction of turbine called 
the Spiro turbine, a phantom 
view of which is shown in Fig. 
99. As will be seen, it con- 
sists essentially of two double 
helical gears rotating in a 
cast-iron casing, which closely 
surrounds them. The steam 
enters through two openings, 
one on each side of a central 

ridge in the bottom of the casing, Fig. 100, and is admitted to a 
tooth space at the center of the rotors. As the gears rotate, this tooth 
space extends in length across the entire width of turbine, thus expand- 




5 team Inlet 



Fig. 100. Section of Spiro Cylinder and 
Rotors at Midlength 



118 STEAM TURBINES 

ing the steam. The maximum ratio of expansion is stated by the 
makers to be 1 to 6. This turbine, therefore, is only run non-condens- 
ing and is useful where steam economy is of secondary importance 
compared with simplicity and compactness. 

REACTION TURBINES 

In this type of turbine, the steam expansion takes place, not 
alone in the nozzles and guide passages, as in all the types of turbines 
previously described, but in the revolving vanes as well, approxi- 
mately half the expansion taking place in each. In this type of tur- 
bine the difference in pressure on the two sides of the vanes brings 
about a leakage of steam over the tips of both the rotating and the 
guide vanes. That this leakage may not be unduly large, the drop 
in pressure is made small from stage to stage. The leakage at the 
high-pressure end gradually expanding does some work in the suc- 
ceeding stages and becomes relatively small at the low-pressure end. 
Clearances, however, in this type of turbine are all-important and, 
other things being equal, that reaction turbine showing the best 
economy will be the one with the smallest radial clearance. To make 
the leakage in the high-pressure stages as small as possible, the vanes 
should be made as long as convenient, about five per cent of the 
diameter of the rotor being considered a minimum. 

In turbines of the Rateau type, there is a diaphragm separating 
the different stages. This extends close to the shaft, permitting 
leakage only through the small annular space between shaft and 
bushing, a comparatively unimportant matter, as these clearances 
may be made very small. Besides this, carbon or labyrinth pack- 
ing may be provided, which will make the leakage from stage to 
stage almost negligible. Moreover, there is no tendency to leak 
over the tips of the running vanes, because the pressure is the same 
on both sides of them. 

In the reaction type of turbine, however, there is no diaphragm 
from stage to stage and, instead of each rotating wheel being a 
separate element, it is customary to fasten the vanes in rows to the 
rotating drum or cylinder. In this style of machine one set of guide 
vanes and one set of revolving vanes constitute a stage, and it can 
readily be seen that the opportunity for leakage between the tips of 
the stationary vanes and the drum of the rotor is very much greater 



120 STEAM TURBINES 

than in the Rateau type, the annular space in the two types being 
proportional to the diameters of shafting and drum. The tendency 
to leak over the tips of the running vanes is even greater, because of 
a greater diameter. However, friction losses, which are approxi- 
mately proportioned to the square of the steam velocity with reference 
to the vanes, will be very much less in the reaction type of turbine 
than in the impulse type, because the steam velocities are compara- 
tively low. 

Parsons Turbine. The Honorable Charles A. Parsons of Eng- 
land is responsible for the successful development of the reaction 
type of turbine. His first turbine, made in 1884, was of 10 h.p., 
18,000 r.p.m., and when running non-condensing with 92 pounds of 
steam pressure by the gage, it is claimed that only 25 pounds of 
steam per brake horsepower per hour were consumed. In 1888, a 
50-h.p. turbine at 7000 r.p.m. was constructed, and soon after a 
200-h.p at 4000 r.p.m. showed good economy. It must be remem- 
bered that the chief problem of the turbine designer has been to 
reduce rotative speeds without material sacrifice in economy. 

Parsons turbines are manufactured in the United States by the 
Westinghouse Machine Company of Pittsburgh, and by the Allis- 
Chalmers Company of Milwaukee, Wisconsin, the former acquiring 
the right to manufacture in 1885, and putting the first turbine on 
the market three years later. For marine purposes, licenses to build 
Parsons turbines have been issued to several firms. 

The essential features of the Parsons turbine are clearly illus- 
trated by the cross section shown in Fig. 101. Steam enters at E 
and, in passing through the annular space between the cylinder walls 
and rotating elements, gradually expands in volume until it exhausts 
at G. The rotor is usually built in three different diameters to 
facilitate mechanical construction and to avoid excessively small and 
excessively large vanes. It is thus possible to use a large number of 
vanes of the same size. When the length of vane would otherwise 
become too great, the same passage area may be provided by short- 
ening the vane and increasing the diameter of the drum. While, 
theoretically, the passage area of the vanes should gradually increase, 
it is found in practice that without any detrimental effect in economy 
several rows of vanes may be made of the same height, and thus the 
areas will increase step by step instead of in a gradual curve. 



STEAM TURBINES 



121 



Since the pressure of steam is greater on the steam side of the 
vanes than on the exhaust side, there will result an end thrust which 
must in some way be balanced. This thrust, due to the static pres- 
sure of the steam, is augmented by the thrust on the vanes due to 
the impact and reaction of the steam in passing through them. To 
balance these thrusts, the balancing pistons shown at L, M, and A 7 , 
Fig. 101, are provided. These are connected to the steam and 
exhaust spaces by the passages 0, P, and Q, so that the pressure can 
be readily balanced. It is, of course, necessary to provide some sort 
of thrust block to meet the requirements of varying conditions, but it 
need not be large, and it serves in general as an adjustment bearing 




tfig. 102. View of Reaction Blading, Showing Method of Mounting 
Courtesy of Westinghouse Machine Company, East Pittsburgh, Pennsylvania 

to keep the rotor in correct alignment, and is so arranged that the 
longitudinal position of the rotor may be slightly changed if desired. 
A mark on this block shows when the rotor is in its correct position. 
Casing. The casing is made with diameters to accommodate the 
various sizes of drums and blading on the rotor. On the inside of 
the casing are the rings of fixed vanes or guides which fit between the 
rings of rotating vanes on the drum. The casing is divided hori- 
zontally, so that by lifting the cover all working parts are exposed. 
In very large turbines the cover slides over four graduated guides, 
one at each corner, so that in lifting it the engineer can readily see 
that it is always kept horizontal while being moved, in order to 
avoid binding or injuring the blading. 



124 



STEAM TURBINES 



Vanes. The vanes are made of suitable material, drawn to the 
proper section, heat treated, and cut to length. The root ends are 
slightly upset and a small hook or shoulder formed, as shown in 
Fig. 102. The grooves in the rotor are of dovetail section and have 
a small auxiliary groove of rectangular section in the bottom. The 
steel distance pieces fit snugly in the dovetailed grooves and the 
shoulders on the lower end of the blades fit down into the rectangular 
groove and hook under the distance pieces. This is a very strong 




<Q! 



e e 



D 








Fig. 105. Bearings of Westinghouse-Parsons Turbine 



construction, as the blades could not be pulled out without actually 
shearing the metal at the shoulder. To stiffen the outer ends of the 
blades, and to maintain a uniform spacing, a wire lacing is threaded 
through openings near their outer edge, twisted between the adjacent 
blades, and soldered into position. 

The Westinghouse Machine Company use a cold-drawn wire 
lacing or lashing of comma-shaped cross section. This is threaded 
through similarly shaped holes near the tips of the blades, and then 



STEAM TURBINES 125 

the tail of the comma is bent over between the blades, holding them 
firmly together, as shown in Fig. 102. 

Rotor. The rotor, as before noted, consists of a drum of three 
diameters. This drum is usually built up of a hollow steel casting 
fixed to some form of spider. The solid rotor would be prohibitively 
heavy. Fig. 103 shows the Westinghouse-Parsons rotor with balance 
pistons. Fig. 104 shows this turbine in cross section. 

Bearing. The rotor being long and heavy, some sort of flexible 
or self-aligning bearing is desirable, if not absolutely necessary. 
Flexibility is provided in the De Laval turbine by means of the long, 
slender shaft, and in Westinghouse turbines, running at more than 
3000 revolutions per minute, by a nest of loosely fitting concentric 
bronze bushings, with sufficient clearance to permit a continuous oil 
film between each two bushings. This is intended to form a cushion, 
permitting a certain amount of vibration in the shaft, yet restraining 
it within very narrow limits. 

It has been found by experience that in the large and slower 
running turbines this so-called flexible bearing is not necessary. For 
these, the Westinghouse Machine Company have now adopted the 
style of bearing shown in Fig. 105. Turbines with short rotating 
shafts would not require a special form of bearing. 

Allis=Chalmers Turbine. The turbine built by the Allis- 
Chalmers Company of Milwaukee is of the Parsons type with a few 
special modifications. Fig. 106 is a longitudinal section through a 
turbine showing all essentials and omitting minor details. Steam 
enters from pipe C after passing through the main regulating valve, 
which is under the control of the governor. Steam enters the cylinder 
through passage E and, turning to the left, passes through the alter- 
nate stationary and moving blades, finally passing to the condenser 
through connection G. The balancing pistons are shown at L, 
My and N. In the large sizes the piston N is placed at the other end 
of the spindle; this reduces the size of the piston and makes it much 
stiff er, as it is backed up by the body of the spindle. The cylinder is 
divided longitudinally into three sections, the end pieces being cast 
separately; it is also divided horizontally in its central plane. 

In the larger diameters the blade rings are made separate 
from the body of the spindle, and are provided with a taper fit and 
pressed on. The balance pistons are constructed in the same way. 



126 



STEAM TURBINES 




fy>m}»)))»»^» 



STEAM TURBINES 127 

The bearings are of the ball-and-socket type with spherical 
seats. The oil is supplied to each of the four bearings by means of 
pipes located in the lower half of the casing. It passes through the 
bearing shells and is admitted to the journal at the middle of the top 
shell. The chief purpose of the collar bearing at the end of the 




Fig. 107. Method of Fastening Blades of 
Allis-Chalmers Turbine 

spindle is to give an accurate axial adjustment. It will also take 
up any small unbalanced thrust which may occur. 

In all reaction turbines, expansion of the casing is a trouble- 
some feature. As this is due to the varying temperatures, there is an 
appreciable difference in the endwise expansion of the spindle and the 
casing. The high-pressure end of the spindle is held by a collar 
bearing, and the difference in expansion is taken up at the low- 



128 



STEAM TURBINES 



pressure end. The labyrinth packing employed at the high-pressure 
end has small axial and much radial clearance, while the labyrinth 




packing of the balance piston at the low-pressure end may have small 
radial clearance but must have large axial clearance to provide for 



STEAM TURBINES 



129 



the difference in expansion. The Allis-Chalmers Company claim 
that this type of construction permits smaller working clearances in 
high-pressure and intermediate pistons. 




Fig. 109. Sets of Blades Assembled 
Courtesy of Allis-Chalmers Company, Milwaukee Wisconsin 

Blading. The general shape of vanes of the Allis-Chalmers 
turbine does not differ from that of other reaction turbines, but the 
method of securing the blade to the casing and to the rotor is different 
from that adopted in the ordinary Parsons type. Each blade is so 
formed that, at its root, it is of an angular, dovetail shape, and has a 



130 STEAM TURBINES 

small projection at its tip. To hold the roots of the blade firmly 
there is a foundation ring A, Fig. 107, which, after being formed to a 
circle of the proper diameter, has slots cut in it by a special milling 
machine, these slots being so shaped as to receive the roots of the 
blade. They are at the same time accurately spaced and so cut as 
to give the required angle to the blades. To protect the tips of the 
blades and to bind them together in a substantial manner, a channel- 
shaped shroud ring B, Fig. 107, is fitted. The small projections on 
the tips of the blade fit through holes cut in this shroud ring and are 
riveted over. These rings, A and B, cover half the circumference and 
the blades are assembled on them, as shown in Fig. 108, before being 
put onto the rotor or casing. These foundation rings are of dove- 
tail shape in cross section, and are inserted into corresponding 
grooves in the turbine casing and spindle, in which they are firmly 
held by key pieces. These key pieces are driven into place and upset, 
so as to fill a small undercut, shown at C in Fig. 107, thus securely 
locking them into place. This construction is applied to all blades 
of whatever length. 

The flanges of the channel-shaped shroud rings are made thin, 
so that, in case of contact with the casing from any accidental cause, 
no dangerous results are likely to follow, the accidental touch merely 
causing a slight wearing away of the flanges of the shroud without 
excessive heating. It is claimed that this shape of shroud ring acts 
in a measure like a labyrinth packing, retarding appreciably the 
leakage of steam. Fig. 109 shows a number of sets of blades as 
assemb'ed. 

COMBINED IMPULSE AND REACTION TURBINES 

A feature of recent development in turbines of large powers has 
been the construction of combined impulse and reaction machines. 
The Westinghouse Machine Company was a pioneer company in the 
development of this type of turbine. The combination consists of 
a Curtis wheel at the high-pressure end and Parsons blading at 
the low-pressure end. The Curtis wheel extracts from 20 to 50 
per cent of the total expansion energy and thus replaces a large 
number of rows of Parsons blading at the point of their lowest 
efficiency. The effect is to greatly shorten the rotor and avoid some 
serious mechanical difficulties which would be met in large power 



STEAM TURBINES 



131 



high-pressure turbines with pure Parsons blading. Fig. 110 shows 
in diagrammatic form a section through a single-flow turbine of this 
class. Steam enters the nozzle chamber and is expanded in the 
nozzles and discharged against a portion of the periphery of the 

Head ion Element ^^^^^ 

-Nozzle Chamber 

■jutse Wheel 




Fig. 110. Section of Combination Impulse and Reaction Single-Flow Turbine 
Courtesy of Westinghoicse Machine Company, East Pittsburgh, Pennsylvania 

impulse wheel. The intermediate- and low-pressure stages are the 
same as those of a pure Parsons turbine. When the steam enters 
the cylinder, its pressure and, therefore, its temperature have been 
much reduced, thus subjecting the cylinder to a small temperature 
difference only. 



F?eaclion. Element 



Impulse V/heel 



Read ion Element 




Exhaust 



Fig. 111. Section of Double-Flow Steam Turbine 
Courtesy of Westinghouse Machine Company, East Pittsburgh, Pennsylvania 

Double=Flow Turbine. The capacity of a turbine depends on 
the weight of steam passed per unit of time; this, in turn, depends 
on the velocity and the height of the blades. For a given rotative 
speed the mean diameter of the blade ring is determined by the 
allowable stress due to centrifugal force, and there is a practical 



132 



STEAM TURBINES 



limit to the height of the blades. In the double-flow turbine, Fig. 
Ill, the capacity is doubled by dividing the current of steam after 
it has passed through the impulse wheel. This type of machine is 



Single Flow Reaction Element 
Reaction Element 



Impulse Wheel 



Dummy 
Reaction Element 




Nozzle Chamber 



Fig. 112. Section of Semi-Double-Flow Steam Turbine 
Courtesy of Westinghouse Machine Company, East Pittsburgh, Pennsylvania 

therefore especially adapted for the largest capacities. An incidental 
advantage of this construction is that the use of dummy or balancing 
pistons is entirely obviated. 




Fig. 113. Rotor of Semi-Double-Flow Steam Turbine 
Courtesy of Westinghouse Machine Company, East Pittsburgh, Pennsylvania 

In the semi-double-flow turbine, Fig. 112, the steam, after 
passing through the impulse wheel, passes through a single section 
of intermediate-stage Parsons blading. It then divides, half of the 
steam flowing through the drum to half of a double-flow low-pressure 



STEAM TURBINES 133 

reaction blading and then into a double-exhaust opening at either 
end of the turbine. This construction is now used by the Westing- 
house Machine Company for alternating-current turbogenerators in 
capacities of from 3000 to 10,000 kilowatts, at speeds of 1500 to 
1800 revolutions per minute. Fig. 113 shows the construction of the 
rotor of a semi-double-flow turbine. Each of these constructions 
has its particular field of usefulness, depending on the operating 
conditions and the speeds for which the turbines are designed. 

GOVERNING 

In order to regulate the supply of steam in proportion to the 
load on the steam turbine, there are several devices employed: (1) 
The pressure of the steam before admission to the steam chest may 
be controlled by means of a throttling governor. (2) In the case of 
impulse turbines with a series of nozzles, each may, if desired, be con- 
nected to an independent valve under the control of the governor, 
so that the number of open nozzles and the extent of the steam belt 
acting on the turbine may be varied to conform to the load, the steam 
being admitted to the nozzles at full pressure. (3) Steam may be 
admitted to the turbine at full pressure at intervals of long or short 
duration. Any of these devices may be arranged to take care of an 
ordinary amount of overload. For very large overloads, a by-pass 
valve is generally provided, which may be controlled either by hand 
or by the governor, admitting high-pressure steam into one of the 
lower stages of the turbine. 

Of these devices, the second evidently is not applicable to the 
reaction turbine or to an impulse turbine taking steam around 
the entire periphery. The other methods of governing might be 
used for any form of turbine. 

Throttling. The simplest governor is undoubtedly the throttle 
type, but throttling is not an economical way of regulating power; for, 
by the throttling process, the steam expands somewhat to a lower 
pressure without doing useful work. The work done goes to super- 
heat the steam and, although no heat units can be destroyed, less can 
be recovered from a pound of the throttled steam than from a pound 
at the initial pressure. The smaller the load, and consequently the 
more the throttling, the greater this loss will be. Any form of tur- 
bine, therefore, taking steam about its entire periphery and fitted 



134 



STEAM TURBINES 



with a throttling governor, would probably be relatively uneconom- 
ical at light loads. 

It is not economical to throttle the steam pressure before admis- 
sion to a set of properly designed nozzles. Nozzles are designed for a 
definite steam pressure and, if this is varied, the efficiency is bound to 




Fig. 114. Sectional View of Governor for Varying Number of Nozzles Opened 

fall off. Consequently, there is a twofold loss due to throttling in a 
turbine of this type. In turbines like the De Laval, supplied with 
several nozzles, any of which may be opened or closed by hand, it can 
readily be seen that if there are, for instance, four nozzles, three of 
which are opened under ordinary loads, then a 33J per cent or even 



STEAM TURBINES 135 

66f per cent under-load may be taken care of by hand regulation. 
A throttling governor on such a turbine would necessarily regulate 
only between these limits and might act only upon one nozzle. On 
account of its simplicity, the throttling type of governor has been 
largely used and is practically always used on the smaller turbines, 
where it has given results that are satisfactory. 

Varying Number of Open Nozzles. Turbines employing partial 
admission through a group of nozzles like those of the Curtis turbine 
may regulate the steam supply by providing an independent valve 
for each nozzle. This valve, being under the control of the governor, 
may be opened or closed as needed. In the Curtis turbine, steam is 
admitted through a series of valves, the number of which depends 
upon the capacity of the machine. The valves are arranged to open 
successively, two-thirds of them being open at full load. The action 
of the valves is so regulated that they are either fully open or fully 
closed. Any increasing load is taken care of by the opening of an 
additional valve, this valve closing when the load falls off. Very 
wide variations in load can thus be carried with little effect on steam 
economy. The friction, of course, is a very much larger per cent 
of small loads, so that the steam consumption at small loads would 
appear very much larger even though it worked with the same 
absolute efficiency as at high loads. 

These valves are controlled from the governor either by means 
of some electrical device, by a direct mechanical control, or by 
hydraulic pressure. The power to operate one of these valves is 
necessarily so great that it cannot be moved directly by the governor, 
but the governor can operate a small pilot valve, which in turn may 
set some mechanism in motion powerful enough to do the work. 

The governor itself is of the centrifugal type, its action depend- 
ing on the balance between the forces exerted by the springs and 
the centrifugal forces of the revolving weights. Fig. 114 shows a 
sectional view of this governor. A is the revolving weight which, 
acting through the knife, edges B, may move the rod C against the 
action of the spring D. At E is a ball-bearing gimbal joint which 
forms a junction between the revolving mass of the governor and 
the stationary lever of the governing arm. The governor is provided 
with the auxiliary spring F, which may be used to vary the speed of 
the turbine. G is a small motor, which may be operated from the 



136 



STEAM TURBINES 



switchboard to regulate the tension of this spring F, thus varying 
the speed of the turbine. The movement of the governor lever is 
transmitted through a connecting rod and lever to the pilot valve 
of an hydraulic cylinder. To provide for overloads greater than 
50 per cent, an auxiliary set of admission nozzles is provided in the 
second stage of the turbine which may take full-pressure steam if 
needed. 

Varying Time of Admission. The third method of governing, 
that by varying the time of admission, is almost invariably used on 




Fig. 115. Turbine Governor for Varying Time of Admission 

turbines of the Parsons type. An admission of steam occurs about 
once in every thirty revolutions at approximately full load. The 
pilot valve is continually oscillating, thus preventing any liability of 
sticking, but its period of oscillating is varied directly by the governor. 
Hence, steam enters the turbine in puffs, the duration of which 
depends upon the load; at slight overloads the valve will be con- 
stantly open. One disadvantage of this method of governing is that 
there may be an opportunity for the turbine to cool between suc- 
cessive blasts of steam, thereby causing initial condensation at 



STEAM TURBINES 137 

moderate and light loads. A by-pass valve, under the direct control 
of the governor, is provided on Parsons turbines. For considerable 
overload this valve will open, letting steam into the intermediate 
stages. 

The details of the mechanism for controlling steam-turbine 
governors are not essentially different from those used in reciprocat- 
ing engines, except such as would be required by the greater speed 
of revolution. In turbines of large size, the valve cannot be oper- 
ated directly by the governor, but must always be moved by means 
of pressure admitted through a small pilot valve. Fig. 115 shows 
the diagrammatic arrangement of the Westinghouse-Parsons govern- 
ing gear. A is the pilot valve under the direct influence of the 
governor which admits steam to the large piston, shown at the left. 
This piston has pressure enough upon it to operate the admis- 
sion valves. 



INDEX. 



INDEX 

A PAGE 

Advantages of steam turbine 3 

Allis-Chalmers reaction turbine 125 

blading 129 

Avery & Foster turbine 7 

B 

Bearings 65, 83, 110, 125 

Blades assembled on rings 128 

Blades, method of fastening 127 

Blades, sets of assembled 129 

Blading 129 

Branca's impulse turbine 5 

Buckets for Curtis turbine 107 

Buckets of Kerr turbine 95 

By-pass valve 83 

C 

Casing. _ 107, 121 

Commercial turbines 55 

impulse 55 

compound 68 

single-stage 55 

reaction 118 

Compound impulse turbines with velocity steps 68 

Compound turbine with one wheel 10 

Compounding 20 

Curtis compound impulse turbine 70, 191 

bearings 110 

casing 107 

diaphragm 109 

guides. 109 

nozzles 107 

vanes 108 

wheels 108 

Curtis turbine and d.c. generator partly assembled 102 

D 

De Laval reaction turbine 10 

De Laval single-stage impulse turbine 56 

bearing 65 

gears 64 

nozzles 57 

shaft 62 

vanes 58 



2 INDEX 

De Laval single-stage impulse turbine (Continued) page 

wheel 60 

Diaphragms 81, 82, 91, 109 

Double guide vanes 70 

E 

Economy curves 49 

Economy of turbine 48 

Elements of steam turbine 1 

Expanding nozzle 20 

F 

Fundamental principles of steam turbine 12 

G 

Governing steam turbine 133 

throttling 133 

varying number of open nozzles 135 

varying time of admission 136 

Guide vanes 109 

H 

Hamilton-Holzwarth compound impulse turbine 88 

Hartman's compound impulse turbine 9 

Hero's steam turbine 5 

History of the steam turbine 5 

I 

Impulse, definition of 15 

Impulse turbines 22 

compound, with pressure stages 78 

Hamilton-Holzwarth 88 

Kerr 95 

Rateau 80 

Wilkinson 93 

Zoelly 84 

compound, with pressure stages and velocity steps 99 

Curtis 101 

Riedler-Stumpf 111 

Terry 112 

compound, with velocity steps 68 

Curtis 70 

Riedler-Stumpf 70 

Sturtevant 74 

Terry 71 

single-stage 55 

De Laval 56 

Riedler-Stumpf 66 



INDEX 3 

PAGE 

Indicated horse-power _. 50 

Installation of steam turbine 41 

J 

Jet and vane, relative positions of 18 

Jet deflection 15 

K 

Kerr compound impulse turbine 95 

Kerr turbine, sectional view of 97 

L 
Low-pressure turbines 32 

M 
Multi-stage turbine 14 

N 
Nozzles 19, 57, 98, 107 

Nozzles and buckets, diagram of . in Curtis steam turbine 100 

Nozzles, wheel, and diaphragm of Kerr turbine 96 

O 

Open nozzles, varying number of 135 

Overhung wheels of Riedler-Stumpf turbine 112 

Overload 48 

P 

Parsons reaction turbine 120 

bearing 125 

casing j 121 

rotor 121 

vanes 125 

Performance of steam turbine 45 

Perrigault & Farcot turbine 9 

R 

Rateau compound impulse turbine 80 

bearings 83 

by-pass valve 83 

casing 82 

diaphragm 81 

vanes 82 

wheels 82 

Reaction, definition of 15 

Reaction turbines 22, 118 

Allis-Chalmers 125 

Parsons 120 



4 INDEX 

PAGE 

Reaction wheel of Avery & Foster 6 

Real & Pichon compound turbine __ 7 

Regenerator __ 39 

Revolving buckets for Curtis steam turbine 109 

Riedler-Stumpf 

compound impulse turbine , 70, 111 

single-stage impulse turbine 66 

Rotating wheels of the Curtis steam turbine 108 

Rotor 122, 125 

Rotor and shaft of Terry single-stage turbine 71 

S 
Shaft __ 62,96 

Single-stage impulse turbines 55 

Single-stage turbine 14 

Speeds of rotation of Curtis turbines 103 

Steam accumulator 39 

Steam consumption of turbine 46 

Steam consumption tests, table 52 

Step bearing of Curtis turbine 111 

Superheated steam , 46 

T 

Table, steam consumption tests 52 

Terry compound impulse turbine 71,112 

Tests of reciprocating engines 50 

Three-bearing support of Rateau turbine 86 

Throttling 133 

Turbine governor for varying number of nozzles opened 134 

Turbine governor for varying time of admission 136 

Turbines 

advantages 3 

compounding 20 

governing 133 

history 5 

impulse 55 

installation 41 

nozzles 19 

performance 45 

reaction 118 

tests 50 

types i 22 

Types of turbines _ .. 22 

V 

Vanes 58, 82, 85, 108, 125 

Velocity of flow 18 

Velocity of whirl ^ 17 

Vertical Curtis turbine of 9000-K. W. capacity 104 



INDEX 5 

W PAGE 

Westinghouse turbo-generator _ 42 

Wheel and vanes of Zoelly turbine 88, 85 

Wheels 60, 82, -88, 96, 108 

of Kerr turbine, assembled on shaft 96 

of Rateau turbine, balanced by holes in disk 84 

of Sturtevant turbine 74 

Wilkinson compound impulse turbine 93 

Wilson's compound turbine 8 

Wolfgang de Kempelen turbine 5 

Working parts of De Laval turbine 63 

Z 

Zoelly compound impulse turbine 84 

vanes r 85 

wheel 88 



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